3--Mechanical Redesign - Pennsylvania State University...Senior Thesis Report Spring 2005 Twin-Tower Enthalpy Recovery Loop, Heat Pipe Heat Exchangers, Fixed Plate Exchangers Thermosyphon

-11 - Florida State University Ringling Conservation Center

Sarasota, FL

Linda Lewis Mechanical Option Senior Thesis Report Spring 2005

3--Mechanical Redesign

Redesign

The goals for The Florida State University Ringling Conservation Center redesign is to reduce

energy consumed by the building as well as provide a means to meet the design conditions

specified by the owner. These design conditions are to maintain the building conditions at 70°F

and 50% relative humidity at all times in areas of the building that will contain art work. The

first redesign is actually an addition to the current building mechanical system. The proposal is

to add energy recovery wheels to all eight of the building air handlers to reduce the cooling coil

load. The wheels will hopefully also be able to help humidify and heat in the winter. The

second redesign is to evaluate the effectiveness of using thermal storage in the campus chiller

plant. Both cases of load leveling and full storage will be evaluated.

Supporting Information for Selected Redesign

Energy Usage

Approximately 90% of the United States energy use comes from Fossil Fuel, mainly oil, natural

gas and coal. In the United States 71% of our electricity is generated using thermal means; i.e.

burning fossil fuels. While this has proven to be an effective way for the United States to

produce electricity, heat and cool buildings, power factories and run vehicles it is becoming

detrimental to our environment. In the process of burning these fossil fuels Nitrogen Oxides

(NOx), Sulfur Oxides (Sox), Carbon Dioxide (CO2), Carbon Monoxide (CO), Volatile Organic

Compounds (VOC) and particulates are produced. These emissions are released into our

environment. They produce many harmful effects to not only our environment but to our health

as well. Acid precipitation results from the solution of nitrogen and sulfur oxides to give a

mixture of nitrous, nitric, sulfurous and sulfuric acids. Carbon Monoxide and Carbon Dioxide

lead to “Greenhouse Effect”. The “Greenhouse Effect” is heating of the environment because

heat loss from the surface of the earth through the atmosphere is reduced by reflection of infrared

radiation from gases and vapors such as CO2 and water vapor. CO and CO2 form CFC’s

(Chlorofluorocarbons) that lead to ozone depletion. The ozone protects us from the harmful

effects of extra ultra-violet radiation from the sun, which ultimately leads to skin cancer and


Sarasota, FL


other health affects such as VOC’s in sunlight react with ozone to produce highly reactive

compounds that attack the human lungs. No one quit knows its full effects yet but changes in our

climate have been documented and contributed to global warming.

According to DOE (Department of Energy) HVAC (Heating Ventilating and Air Conditioning)

Systems in commercial and residential buildings accounts for 40-60% of energy consumption in

the United States. The graphs below come from Energy Consumption Characteristics of

Commercial Building HVAC Systems written by Detlef Westphalen and Scott Koszalinski for the

DOE. From Figure 3.1 it can be seen that the south uses the most energy of any U.S. region and

about 83% of the energy used by the south is consumed by HVAC.

Figure 3.1 Energy Usages in the United States

It has taken years to get our energy production and consumption to the point at which they are

now, it will also take years to turn it back around. There are many ways to reduce energy

consumption though. One of the main ways is by using renewable recourses to produce energy

(wind energy, geothermal energy, Solar, Hydro-electric and biomass). HVAC loads can be

reduced by using higher efficiency equipment and/or less energy consuming equipment. This

would include but is not limited too: low e glass, tighter building envelopes, energy reduced


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equipment (such as Energy Star rated), energy reduced light fixtures, day lighting, water

reducing fixtures, economizers, heat recovery systems, and cogeneration,

The Florida State University Ringling Conservation Center is located in Sarasota Florida.

Reduction in energy usage in this area is very important to our environment as well as our health.

Florida Power and Light will supply the Florida State University Ringling Conservation Center.

They offer some incentives for reduces electricity consumption:

• Rebates are available for energy efficient equipment including lighting, air conditioning,

chillers, thermal storage, insulation and window treatments, and other custom measures.

• Free Business Energy Evaluations are also offered, providing comprehensive analysis of

facility energy use and recommendations for cost-effective energy efficiency

improvement.

• The Business Custom Incentive Program rewards innovations that trim at least 25 kW

from peak demand. To qualify for a business custom incentive, the project must differ

from other FPL conservation programs and pass the FPL’s energy conservation test.

Indoor Air Quality (IAQ)

In general terms IAQ is how the indoor air affects the health and well-being of those exposed to

it. In more technical terms IAQ is defined by how indoor air satisfies three basic requirements

for human occupancy:

1. Thermal acceptability.

2. Maintenance of normal concentrations of respiratory gases.

3. Dilution and removal of contaminants and pollutants to levels below health or odor discomfort thresholds.

IAQ is not a simple nor is it easily defined. It is a constantly changing interaction of complex

factors. These factors affect the types, levels, and importance of pollutants in indoor

environments. They include: sources of pollutants or odors; design, maintenance and operation

of building ventilation systems; moisture and humidity; and occupant perceptions and


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susceptibilities. In addition, there are many other psychological factors that affect comfort or

perception of IAQ.

Some of the issues with IAQ today are:

• The Environmental Protection Agency (EPA) has estimated that we spend 90% of our

time indoors • Statistics show that 1 out of 5 Americans suffer from allergies caused by substances

found in the home and office.

• Asthma-

o In adults, asthma can develop in response to irritants in the workplace -

chemicals, dusts, gases, moulds and pollens.

o Nearly one in 13 school-aged children has asthma

o Asthma is the leading cause of school absenteeism due to a chronic illness,

accounting for over 14 million missed school days per year

o Deaths related to asthma have risen 40% in the past two decades.

• Indoor Air Quality problems can result in discomfort, acute health effects, and/or chronic

health effects for the building occupants

• Mild effects of poor IAQ are: irritation of the eyes, nose, and throat, headaches,

dizziness, and/or fatigue

• Serious effects of poor IAQ: respiratory cancer, chronic obstructive pulmonary disease or

immunologic disorders

• Poor indoor air quality can result in a decrease in productivity of workers and larger

numbers of absenteeism’s.

NOTE: The health effects associated with indoor air pollutants is difficult to access since

documented cases are hard to come by. Health effects may show up immediately or years later as

a result of pollutants. Health effects associated with chronic low-level exposure to common

indoor air pollutants still remain unexplored.


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Sick Building Syndrome (SBS) has been an increasing research topic in recent years. SBS can be

defined as a persistent set of symptoms occurring in greater than 20% of those exposed, with

cause(s) not recognizable, and complaints/symptoms relieved after exiting the building.

Diagnosis of SBS is primarily based upon the exclusion of other disease and is determined

essentially by perception. These symptoms include but are not limited to: eye, nose or throat

irritation, headaches, fatigues and dizziness, difficulty in concentration, irritability, Dry skin,

rashes, Nasal congestion, difficulty in breathing, nose bleeds and nausea. These symptoms lead

to what is known as building related illness (BRI). Some BRI’s are hypersensitivity pneumonitis,

pontiac fever, legionnaire's disease and humidifier fever. The World Health Organization

(WHO) estimates that 30% of all commercial buildings exhibit signs of "sick" tendencies.

Poor IAQ can be contributed to the pollutants inside a building. Pollutants inside the building

can come from:

• Outside- cars, trucks, etc.

• Electrical equipment- computers, copy machines, printers, etc.

• Cleaning supplies- floor wax, carpet deodorizers, air fresheners ,etc.

• Cigarette smoke

• Off Gassing - new carpeting, furnishings, etc.

• Insulation and window coverings

• Pressed wood products

• Microbes such as mold and fungi – often mold develops in standing water within

the mechanical system and is then distributed throughout the building in the

supply air

• Etc.!!

Without proper ventilation these pollutants make it into our lungs instead of returned to the

mechanical equipment and exhausted or filtered.


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One of the main causes of the increase in building pollutants over the years is the change in

building design and operation. In the 1970’s energy conservation attempts “tighten” buildings

designs. Buildings do not breathe (infiltration and exfiltration) like they used to. Another attempt

at conserving energy is made by owners. Building owners will partially or fully close outdoor air

dampers to reduce energy consumption. They have also been cutting costs on needed

maintenance to mechanical equipment, thus equipment is not being changed or cleaned

adequately. This results in inadequate ventilation for building occupants.

Another cause of poor indoor air quality is using a building for other purposes than it was

designed to be used for. Often there are too many people in an area, resulting in the inability of

the ventilation system to be affective in this area. Also, the addition of electrical equipment

and/or furniture in an area can add to the pollutants and load in the room not allowing for proper

ventilation.

Problems arise when ventilation systems, in an effort to save energy, do not bring in adequate

amounts of outdoor air. The ventilation requirements set by ASHRAE Standard 62 require a

certain amount of outdoor air to be brought into the building at all times. Unfortunately this does

not always occur. Inadequate ventilation can occur if the air supply and return vents within each

room are blocked or placed in such a way that outdoor air does not actually reach certain areas of

the building. Improperly located outdoor air intake vents can also bring in air contaminated with

automobile and truck exhaust, boiler emissions, fumes from dumpsters, or air vented from

restrooms. Finally, ventilation systems can be a source of indoor pollution themselves by

spreading biological contaminants that have multiplied in cooling towers, humidifiers,

dehumidifiers, air conditioners or the inside surfaces of ventilation duct work.

In a cooling application if re-circulated was fully used the load on the cooling coil would only

result from cooling the return air at approximately 75°F instead of cooling the outdoor air that

may be over 100°F. The outdoor air load is thus one of the largest energy consumers in the

building. Ways to reduce the outdoor air load of a building without reducing indoor air quality

are by using desiccant wheels, enthalpy wheels, sensible wheels, Coil Energy Recovery Loop,


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Twin-Tower Enthalpy Recovery Loop, Heat Pipe Heat Exchangers, Fixed Plate Exchangers

Thermosyphon Heat Exchangers and better designs. Better designs are those that decrease

ventilation air required in a building. An example is a dedicated outdoor air system (DOAS) that

only supplies the amount of air (air at the right temperature and absolute humidity to take care of

the latent load in each space) needed to each space to meet its ventilation requirements and uses

another parallel system to take care of the sensible load in the room. Another example is a

variable air volume (VAV) designed with all rooms on the same air handling unit having the

some Z factor (Z factor defined by ASHRAE Standard 62)


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3.1--Energy Recovery Wheels

Background Information

Energy Recovery Wheels can be used in systems to reduce the outdoor air cooling load and

improve the indoor air quality. As mentioned previously the outdoor air load is one of the

largest loads associated with a building HVAC system. Energy wheels run between an exhaust

air stream and the outdoor air stream. The wheels are used to exchange the properties of the

exhaust air stream and with that of the outdoor air stream without actually mixing the two air

streams. Energy Recovery Wheels can also reduce condensation on equipment and humidity

levels in air ducts, eliminating growth of mold, mildew, and bacteria, and thus helping control

indoor air quality. Also, Energy Recovery Wheels inadvertently increase IAQ when they

decrease the outdoor air load because building operators are much less likely to close off or

reduce the outdoor air damper when the energy cost is less. There are three main types of energy

recovery wheels; sensible wheels, enthalpy wheels and desiccant wheels. Sensible wheels

exchange sensible heat only between the two air streams. Enthalpy wheels exchange latent and

sensible heat between the two air streams. Desiccant wheels primary function is to transfer

latent heat, in the process sensible heat is actually added to the air.

The advantages to using Energy Recovery Wheels are:

• Pre-conditioning incoming outdoor air thus reducing the outdoor air load on the system.

• Easily integrated/retrofitted into new/existing ventilation systems.

• Helps to meet ventilation standard without raising energy cost.

• Allows reduction in system capacity by 30 to 65%.

• Less dehumidification and humidification is required

The disadvantages/drawbacks of using Energy Recovery Wheels are:

• Possible cross contamination between air streams


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• Generally not cost effective when the exhaust air stream is less than 50% of the outdoor

air stream.

• Increased maintenance- cleaning wheels

Desiccant Dehumidification

Background

Desiccant dehumidification is the process that uses a desiccant material to produce a

dehumidification affect. The process involves exposing the material to a high relative humidity

air stream where it will collect moisture from the air stream. Conventional solid desiccants

include silica gel, activated alumina, lithium chloride salt, molecular sieves and a new sold

desiccant material is zeolites.. Zeolites are designed to be more effective for cooling

applications. Liquid desiccants include lithium chloride, lithium bromide, calcium chloride, and

triethylene glycol solutions.

After the material is exposed to a high relative humidity air stream, it is then exposed to a lower

relative humidity air stream which will draw the moisture back out. The first air stream is

dehumidified while the second air stream is used to regenerate the desiccant material. In the

process of removing moisture from the air stream the desiccant wheel can also remove

contaminants thus improving IAQ. Often a normal exhaust air stream is unable to regenerate the

wheel to the proper conditions in which it can then remove the amount of moisture needed from

the high relative humidity air stream. A high heat source is needed to regenerate the wheel. This

heat source can be any means that reaches the appropriate temperature for proper regeneration

but one that is low cost or free heat is the best. Examples of ways to obtain regeneration heat are

solar energy, electric heat and waste heat off a boiler or absorption chiller. The process describe

above refers to a passive desiccant wheel, a system with only a desiccant wheel, it can be seen in

Figure 3.1-1 below.


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Figure 3.1-1 Desiccant Dehumidification

An Active Desiccant wheel is another configuration of a desiccant dehumidification system. An

example of one way of configuring an active desiccant system can be seen in Figure 3.1-2 below.

This desiccant system consists of a desiccant, and a heat exchanger often an enthalpy wheel.

The enthalpy wheel will rotate at a speed of about 40 revolutions per minute (rpm) between the

exhaust air from the building and the process air stream. The process air stream as it is supply

side air stream. The desiccant wheel will rotate at a slower speed of about 10 revolutions per

hour (rph). The desiccant is equipped with a burner to modulate the regeneration temperature

and thus control the moisture removal capacity of the wheel.

Figure 3.1-2 Active Desiccant Dehumidification


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Figure 3.1-3 Basic Air Desiccation Process

In Figure 3.1-4 shows the basic air process of a desiccant system. The desiccant wheel is used to

drive the moisture out of the air (process A to B in Figure 3.1-3). The use of a heat exchanger is

used to cool the air back down after the moisture is driven out (Process B to C in Figure 3.1-3).

The cooling down of the air after the dehumidification process and the needed heat to regenerate

the wheel add complexities to the desiccant system that sometimes make the system unable to be

used in certain applications. Normally desiccant dehumidification is best in applications where

the regeneration heat can be taken from a “free” source (like an absorption chiller) and the

cooling of the air after the dehumidification process is also “free” or at a lower energy cost than

it would be to just let the cooling coil do the process, as is normal practice in most conventional

systems today.

Redesign using Desiccant Dehumidification

Calculations

The Florida State University Ringling Conservation Center’s location and parameters make it a

good application for desiccant dehumidification. The humidity levels in Florida reach around

147.5 grains per pound for decent periods of time. Also, the outdoor air temperature in the


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summer months is in the upper 90’s most days. The parameters for the building are that in area’s

with containing art work the relative humidly must be kept at 50% at all times. Using desiccant

dehumidification will help properly maintained relative humidity while lowering the energy cost

of the building systems. The proposal was to using an active desiccant system in each of the

eight air handling units in the building. The configuration of the wheels is shown below in figure

3.1-4. This is like Figure 3.1-3 using the enthalpy wheel as the heat exchanger. Do to the fact

there is no waste heat in the building that can be used to regenerate the wheel an electric heater

must also be installed to create the heat for the wheel regeneration. The exhaust air stream must

be split to provide air for both the desiccant wheel and the enthalpy wheel. Calculations where

done using 285°F and 320°F for the electric heater temperatures. These temperatures are the

limits that Novel Aire suggests for low and high temperatures.

Figure 3.1-4 Active Desiccant System for Redesign

Design Considerations/Assumptions:

• The room air conditions for general occupancy area’s is 72°F and 50% RH

• The room air conditions for area’s containing artwork is 70°F and 50% RH

• A 2°F rise in exhaust air was assumed from room conditions.


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o The exhaust air flow was split between the desiccant wheel and the enthalpy

wheel. Using the Novel Aire desiccant wheel simulation program the amount of

air needed to pass through the desiccant wheel was determined. The rest of the

exhaust air was then used for the enthalpy wheel.

• The effectiveness of the enthalpy wheel and the desiccant wheel where taken from Novel

Aire simulation program. (Shown in Figure 3.1-5 & 6 below for AHU 1-1)

Figure 3.1-5 Novel Aire Desiccant Wheel Simulation Program


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Figure 3.1-6 Novel Aire Enthalpy Wheel Simulation Program

NOTE: Wheel Effectiveness typical = (143.4-111.1)/(143.4-72) = 0.452 = 45.2%

Conclusions Desiccant Dehumidification

The calculations for this section can be seen in Appendix A. After calculating the results of

using this configuration for AHU1-1 it was concluded that this set up of desiccant

dehumidification was not worth using. The savings was compared to the same system just using

an enthalpy wheel. The savings using the enthalpy wheel was greater (this is will discussed

more in the next section). The savings created by using this set up was minimal and the extra

equipment necessary was too great for these little savings. As it can be seen in the results there

are many things that led to a low energy savings in this set up:

• The exhaust air had to be split between two sections thus reducing the efficiency of both

the desiccant wheel and the enthalpy wheel.


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• Running the electric heater add energy to the system (these numbers were not calculated

since the energy savings of the wheel was so low—it was unnecessary to calculated this

additional energy cost.)

• Using a desiccant wheel before an enthalpy wheel reduces the effectiveness of the

enthalpy wheel.

• The sensible load increase from the desiccant wheel is too great for the enthalpy wheel to

handle and thus reduces the effectiveness of this system.

In conclusion an active desiccant system analyzed here was not a good energy reducer. Instead

of this desiccant system enthalpy wheels alone will next be used to reduce the outdoor air load.

Just using an enthalpy wheel is less maintenance as well as more effective. Another active

desiccant dehumidification configuration may have been more effective in this application. All

desiccant dehumidification systems are fairly complicated and add additional maintenance.

Therefore, active desiccant systems will not be used for this building.

Enthalpy Wheels

Background

Enthalpy wheels are becoming common in Heating, Ventilating and Air Conditioning (HVAC)

systems. They bring incoming air closer in temperature and humidity to exhaust air, which

reduces the load on the heating and cooling systems. A well-designed enthalpy wheel system

(one in which the exhaust air flow matches the outdoor air flow) can recover 60% to 80% of the

energy that would otherwise be needed to heat or cool outside air. It lowers building operating

costs and the capital cost of cooling and heating equipment as smaller devices can be installed.

As the same with the desiccant wheel the enthalpy wheel can remove contaminants for the air

thus improving IAQ.

An enthalpy wheel is positioned so that the exhaust air and the supply air travel through it in

opposite directions through separate ducts. The incoming air travels through the wheel before it

enters the rest of the HVAC system. Enthalpy wheels can recover both sensible and latent


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energy. Sensible heat is transferred as the metallic substrate picks up and stores heat from the

warmer air stream and gives it up to the cooler one. Latent heat is transferred as the synthesized

metallic substrate condenses moisture from the air stream that has the higher humidity ratio

through adsorption and releases the moisture through evaporation into the air stream that has the

lower humidity ratio. This all can be seen in figure 3.1-7 below.

Figure 3.1-7 Enthalpy Wheel

Wheels with a honeycomb matrix were introduced in the mid sixties. The medium was asbestos

paper impregnated with lithium chloride. Due to inherent absorption properties of asbestos and

lithium chloride these rotors had a short life. In the late seventies asbestos was replaced by kraft

paper; however, lithium chloride continued to remain the preferred desiccant due to its ease of

impregnation of media. In the mid seventies, two new enthalpy wheel models hit the market and

continue to be offered to date. The oxidized aluminum wheels offered by some manufacturers,

has corrugated aluminum foil wound on a mandrel and braced by steel strips on the sides. The

assembly is dipped into a bromide solution to cause the aluminums to oxidize and form a layer

of alumina - a known desiccant. Such wheels have heat transfer characteristics comparable to the

others at a lower cost. However, these wheels have a weaker structural integrity and suffer from


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a desiccant migration problem. The other type of wheel uses silica gel as desiccant which is

bonded to the aluminum substrate through a coating process. The matrix is supported by spokes

and rim assembly. In the 1980s, considerable advances were being made in the fabrication of

silica and other compounds for the semiconductor industry. A derivative of these innovations

was the development of molecular sieves – synthetic zeolites that could be designed at the

molecular level. At the same time, manufacturing processes had been developed to allow the

bonding of a breathable layer of desiccant to metal or plastic surfaces.

Enthalpy wheels today are mainly either use Silica Gel or Molecular Sieve as the transfer

medium. Silica gel is a highly porous solid adsorbent material that structurally resembles a rigid

sponge. It has a very large internal surface composed of myriad microscopic cavities and a vast

system of capillary channels that provide pathways connecting the internal microscopic cavities

to the outside surface of the sponge. Silica gel enthalpy wheels transfer water by rotating

between two air streams of different vapor pressures. The vapor pressure differential drives

water molecules into/from these cavities to transfer moisture from the more humid air stream to

the drier air stream. Silica Gel is a substance that has preference for the adsorption of water

vapor molecules over other chemicals. Silica gel is the best desiccant for comfort ventilation

applications because, at typical relative humidity’s, it transfers two to three times as much water

by weight as compared to a Molecular Sieve. Silica gel has superior characteristics for

recovering space conditioning energy from exhaust air and handling high relative humidity

outside conditions. Another key point is that the transfer of water by sorption/desorption is not

dependent on temperature. Thus, the silica gel enthalpy wheel works to reduce latent load at

difficult part-load conditions. Molecular sieves are crystalline metal aluminosilicates having a

three dimensional interconnecting network of silica and alumina tetrahedra. Natural water of

hydration is removed from this network by heating to produce uniform cavities which selectively

adsorb molecules of a specific size. Molecular sieves are preferred for regenerated applications

such as desiccant cooling and dehumidification systems that must reduce processed air streams

to very low relative humidity.


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Figure 3.1-8 Enthalpy Wheel Desiccants

Early in the development of Enthalpy wheels cross contamination was a large problem. Faulty

seals, poor designs and poor installations gave Enthalpy wheels a bad reputation in the design

world. They were not readily used after this occurred. Today they are slowly making a

comeback. Early in the development of Enthalpy wheels cross contamination was a large

problem. Faulty seals, poor designs and poor installations gave Enthalpy wheels a bad

reputation in the design world. They were not readily used after this occurred. Today they are

slowly making a comeback. The concern for enthalpy wheels is that the exhaust air could

contaminate the fresh air coming from outdoors with particles that are removed from the

outgoing airstreams by the wheel, such as VOCs (Volatile Organic Compounds) from cleaning

agents, or microorganisms from carpets, behind insulation and even from people. Through the

use of better seals, purge sector and molecular sieve; cross-contamination is reduced to an

insignificant amount, around 0.04%.


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Redesign using Enthalpy Wheels:

Calculations

On The Florida State University Ringling Conservation Center Enthalpy wheels where added to

each of the air handling units to analysis the energy savings that they would achieve. The hope

is to reduce the outdoor air latent load significantly and that the first cost of the wheels is offset

by the savings that could be achieved by adding the wheels to all of the eight units.

Design Considerations/Assumptions:

• The room air conditions for general occupancy area’s is 72°F and 50% RH

• The room air conditions for area’s containing artwork is 70°F and 50% RH

• A 2°F rise in exhaust air was assumed from room conditions.

• The effectiveness of the enthalpy wheels where taken from Novel Aire simulation

program. (Shown in Figure 3.1-9 below for AHU 1-1)

• All of the air handling units where designed to supply air at 48°F and 38 Gr/lb

• The wheels are run until there is no net savings for running them.

• The wheels are modulated to reach design supply air temperatures when the temperature

falls below 48°F (the supply air temperature). While heating the air the wheels also

humidify, reducing the load on the humidifiers. At these conditions the cooling coil load

is not in use.


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Figure 3.1-9 Novel Aire Enthalpy Wheel Simulation Program AHU 1-1

NOTE: Wheel Effectiveness Typical = (147.5-85.7)/(147.5-58.5) = 0.694 = 69.4%

A bin year was created using Bin Maker. It takes in to account occupied/unoccupied times as

well as considers the fact that from May to September the building has a reduced occupancy in

certain areas of the building. This reduction in building occupancy comes from the fact that the

building is an education facility. From the HAP load analysis the average outdoor air flow in

each bin was determined. The total savings for each air handling unit was then calculated and

related to the amount of kilowatts of energy that would be saved. Lastly, a cost analysis was

done to determine the payback period of adding the wheels.

Conclusion--Enthalpy Wheels

Calculations for this section can be seen in Appendix B. Overall the enthalpy wheels proved to

be a sufficient way to reduce the load on the cooling coils of each air handling unit. Table 3.1-1


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shows the Novel Aire wheels that were selected for each air handling unit. Novel Aire’s

specification booklet can be seen in Appendix B.

AHU SA cfm OA cfm EA cfm Novel Aire

ECW Purge cfm Wheel Size

(inches) AHU 1-1 990 590 470 244 4 24 AHU 1-2 2500 1540 1380 364 13 36 AHU 1-3 4870 4870 4870 544 36 60 AHU 1-4 1450 580 480 244 4 24 AHU 1-5 3100 420 320 204 2 20 AHU 2-1 5600 2000 1380 364 13 36 AHU 2-2 1000 7600 4500 724 35 72 AHU 2-3 880 580 480 244 4 24

Table 3.1-1 General AHU Information and Enthalpy Wheel Selections

A summary of the savings achieved by each air handling unit can be seen in Table 3.1-2. As it

can be seen from these results on average the enthalpy wheels saved 33% of the overall outdoor

air load.

AHU Without EW kBtu

With EW kBtu

Savings Total kBtu % Savings

AHU 1-1 240,855 157,954 82,901 34% AHU 1-2 458,110 317,167 140,943 31% AHU 1-3 823,124 533,994 289,131 35% AHU 1-4 197,700 133,077 64,622 33% AHU 1-5 129,219 88,252 40,966 32% AHU 2-1 639,188 435,785 203,404 32% AHU 2-2 1,208,102 855,603 352,498 29% AHU 2-3 109,287 67,995 41,292 38%

Table 3.1-2 Total Outdoor Air Load

The enthalpy wheels were installed mainly to lessen the latent load of the system. A summary of

the latent load savings can be seen in Table 3.1-3 below. The enthalpy wheels were able to on

average save 49% of the overall latent load on the coiling coil. Of the total savings by the

enthalpy wheel on average 84% was latent load. This is because the latent load, especially in


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Florida, is much higher than the sensible load, therefore resulting in a great difference between

the outdoor air stream gr/lb and the exhaust air stream gr/lb.

AHU Without EW kBtu

With EW kBtu

Savings Total kBtu

% Savings of Latent

Load % Savings of Total Savings

AHU 1-1 139,873 70,321 69,552 50% 84% AHU 1-2 259,975 136,647 123,328 47% 88% AHU 1-3 461,681 220,810 240,871 52% 83% AHU 1-4 110,961 57,217 53,744 48% 83% AHU 1-5 74,417 39,201 35,216 47% 86% AHU 2-1 359,788 189,633 170,155 47% 84% AHU 2-2 687,560 399,524 288,036 42% 82% AHU 2-3 64,438 29,493 34,945 54% 85%

Table 3.1-3 Latent Savings

Table 3.1-4 shows the sensible savings. On average the wheels save 12% of the overall sensible

coiling cool load can be reduced using enthalpy wheels.

AHU Without EW kBtu

With EW kBtu

Savings Total kBtu

% Savings of Sensible

Load % Savings of Total Savings

AHU 1-1 100,982 87,633 13,349 13% 16% AHU 1-2 198,135 180,520 17,615 9% 12% AHU 1-3 361,443 313,184 48,259 13% 17% AHU 1-4 86,738 75,860 10,878 13% 17% AHU 1-5 54,802 49,052 5,750 10% 14% AHU 2-1 279,400 246,152 33,248 12% 16% AHU 2-2 520,542 456,080 64,463 12% 18% AHU 2-3 44,850 38,502 6,348 14% 15%

Table 3.1-4 Sensible Savings

The humidifier load in the building is relatively low. The only air handlers that contain

humidifiers are AHU 1-1, AHU 1-3, AHU 1-4, and AHU 2-3. These air handling units serve the

areas where artwork is located. The humidifier load is very hard to determine. It depends really

on the return air conditions. If the absolute humidity in the return air is enough to, when mixed,

offset the dryness of the outdoor air the humidifier is not needed. If the absolute humidity of the


Sarasota, FL


return air mixed with the outdoor air is less than the needed air humidity level then the

humidifier would be needed. The enthalpy wheels can aid in the process of humidifying but can

not always take the full load. These results assume that the enthalpy wheels are running to heat

the air up to supply air conditions. The humidifying that the wheel does in the process of heating

this air up is summarized in Table 3.1-5 and the heating in Table 3.1-6. The air may not actually

need to be heated to the supply air temperature though. Mixing the outdoor air with the return

air may still result in cooling during these bins. Therefore these results are only good when the

outdoor air needs to be heated by the enthalpy wheel to the supply air temperature. It is very

hard to know how much savings the wheels will provide for heating and humidifying. The only

concrete result is that they will provide some reduction in load.

AHU Without EW kBtu

With EW kBtu

Savings Total kBtu

% Savings of Total Savings

AHU 1-1 1,833 1,593 240 13% AHU 1-2 4,093 3,550 543 13% AHU 1-3 6,967 5,921 1,046 15% AHU 1-4 1,900 1,658 242 13% AHU 1-5 1,185 958 227 19% AHU 2-1 5,775 4,941 834 14% AHU 2-2 21,973 20,779 1,194 5% AHU 2-3 434 378 56 13%

Table 3.1-5 Humidifier Load

AHU

Heating Load Btu

Heating Load w/EW Btu

Savings Total kBtu

% Savings of Total Savings

AHU 1-1 242 0 242 100% AHU 1-2 593 0 593 100% AHU 1-3 1,138 0 1,138 100% AHU 1-4 338 0 338 100% AHU 1-5 243 0 242 100% AHU 2-1 1,173 0 1,173 100% AHU 2-2 1,267 0 1,267 100% AHU 2-3 57 0 57 100%

Table 3.1-6 Heating Load


Sarasota, FL


Cost Savings—Enthalpy Wheels

The cost savings of the enthalpy wheels was calculated and is summarized in Table 3.1-7. The

cost savings was done using 36% of the hours as on-peak and 64% of the hours as off-peak.

This is a fair assumption using Florida Power and Lights on-peak and off-peak hours. The

amount of savings achieved by the enthalpy wheels is on average $3,275.68.

AHU On Peak kWh

per Year

Off Peak kWh Per

Year On Peak kWh

Cost Off Peak kWh

Cost AHU 1-1 5,033.7 8948.9 $120.00 $66.31 AHU 1-2 10,667.7 18964.8 $254.32 $140.53 AHU 1-3 21,702.3 38581.9 $517.38 $285.89 AHU 1-4 3,023.3 5374.7 $72.07 $39.83 AHU 1-5 3,087.4 5488.7 $73.60 $40.67 AHU 2-1 15,328.0 27249.7 $365.42 $201.92 AHU 2-2 26,595.1 47280.1 $634.03 $350.35 AHU 2-3 3,062.6 5444.6 $73.01 $40.34

Totals 88,500.1 157333.5 $2,109.84 $1,165.84 Table 3.1-7 Total Cost Savings

The first cost of the enthalpy wheels is shown in Table 3.1-8. These prices are from Novel Aire.

Wheel Price Number of

Wheels Installation

Cost Total Cost ECW 204 $600.00 1 $420.00 $1,020.00 ECW 244 $900.00 3 $420.00 $3,120.00 ECW 364 $1,500.00 2 $505.00 $3,505.00 ECW 544 $3,400.00 1 $630.00 $4,030.00 ECW 724 $6,100.00 1 $630.00 $6,730.00

Totals $12,500.00 8 $2,605.00 $18,405.00 Table 3.1-8 Enthalpy Wheel First Cost

The cost of the wheels is not the only addition cost associated with adding enthalpy wheels to the

system. There is also additional cost for installing more exhaust fans since there are a few larger

exhaust fans in the building. Smaller exhausts fans were added to each air handling unit to

provide the wheels with there appropriate exhaust air cfm. Also the section for the enthalpy

wheel must be added to the unit. The total of all these costs was on average $625 per air


Sarasota, FL


handling unit. Therefore the actual total cost of adding enthalpy wheels to the system is

$24,405. The cost of the coiling coils is also reduced for each unit. This results in a first cost

reduction of $7896.00 therefore reducing the first cost of the system to $15,509.00. The pay

back period for the wheels is approximately 4.7 years.


Sarasota, FL


3.2--Thermal Storage:

Thermal storage is the process of storing energy. The energy can be stored in the storage

medium either by changing its temperature (heating or cooling) or by changing its "phase"

(liquid to solid). The energy is released when the process is reversed. This section will only

focus on storing cooling since that is the application that will be applied to The Florida State

University Ringling Conservation Center.

Long-term storage of thermal energy in large amounts (>1000 MWh) has been in use since the

1980s. The storage medium is generally water or rocks, e.g. in aquifers, rock caverns, bore

holes. Short-term storage of "coolness" in air-conditioned buildings is mostly used for Demand

Side Management (DSM) purposes, where the systems have been found to be cost-effective.

The main applications have been in the USA, Canada and Japan.

The concept of cooling thermal storage is to produce cooling by making ice or low temperature

water at off-peak electrical times (normally night time) to be melted or used during on-peak

electrical times. This is done because off-peak electrical rates are lower than on-peak rates. As

well as on-peak demand charges can be greatly reduced when the peak load is shifted from on-

peak hours to off-peak hours.

Storage can be done as “full storage” or “partial storage”. Installing a system that is capable of

avoiding all the on-peak chiller operation is referred to as “full storage.” An alternative referred

to as "partial storage" minimizes or eliminates any additional initial capital investment. By

operating a chiller for the entire day, on-peak at standard conditions and off-peak at ice-making

conditions, its size is usually reduced to 40% to 50% of the conventional design. A graph of

“full storage” and “partial storage” can be seen below. For “partial storage” the cooling

equipment (a chiller) would not be able to fully shut-down during the on-peak hours. Often

“partial storage” is used to load level (as seen in figure), letting the chillers run at full capacity

all the time. This allows for the best efficiency in your chiller (more energy efficient) as well as

provides a level load profile. A level load can lead to a good application for a combined heat

and power system (CHP) that allows for even more additional energy savings.


Sarasota, FL


Figure 3.2-1 Full Storage Vs. Partial Storage

The HVAC industry is very sensitive to the changes in electric power rates. Thermal storage for

cooling applications depends almost entirely on time differentiated utility rate structures for its

existence. Thermal storage is used mainly in commercial buildings because commercial


Sarasota, FL


building on-peak electrical rates are one of the highest rates out there. The commercial customer

presents a particularly poor load profile to the utility. Therefore, commercial customers pay, on

average, 65% more for their electricity than industrial customers. Considerable debate exists

about the eventual savings that deregulation will produce, but it is clear that off-peak power will

be extremely inexpensive.

There are many advantages to using thermal storage, some of which have been listed above. The

greatest benefit of thermal storage is its ability to produce more kWh from fewer kW of

operating capacity. Producing energy at off-peak times reduces the load on the utilities thus

taking some of the peak load off of them. All other things being equal, thermal storage

customers will consume approximately the same kWh as their conventional system counterpart.

Even if customers can only displace 20% of the peak demand, a conservative goal, the power

provider has the opportunity to sell all of the original kWh, plus an additional 20% in kWh sales

to another customer. Although it is debatable, most of the time Thermal storage systems can be

designed to use less electrical energy than their conventional counterparts. This savings comes

from using equipment at maximum capacity for longer time periods. Equipment sizes are most

often smaller, up to 40-50%. Often in full or partial storage the capacity of the chiller needed is

less than in a non-thermal storage application. Other benefits from this include the reduction in

operating costs compared to on-peak marginal capacity, lower emissions costs per kWh and

improved transmission efficiency. Thermal storage also presents one of the only ways of

shifting load. Storage systems do not negatively impact a facility's operation, as other load

shedding or load control programs almost always do. On a first cost basis thermal storage

systems are typically little to no additional cost and provide an energy cost reduction. Finally,

thermal storage is very versatile. As electrical rates change and costs shift thermal storage can

be adjusted to meet these changes.

As with every system there are draw-backs to using thermal storage. Storage tanks must be

installed. They can often take up large areas and cause maintenance issues. There are also

safety and hygiene concerns when there are large tanks with standing water or ice in them. These


Sarasota, FL


systems are not limited to but are most often used with only pure water and not mixture solutions

for these reasons. Special training for operators and maintenance persons can also be a concern.

Cool thermal storage can be done with chilled water or ice as the storage medium. Chilled-water

storage systems use the sensible heat capacity of water—1 Btu per pound (lb) per degree

Fahrenheit (F)—to store cooling capacity. They operate at temperature ranges compatible with

standard chiller systems and are most economical for systems greater than 2,000 ton-hours in

capacity. Ice thermal storage systems use the latent heat of fusion of water—144 Btu/lb—to

store cooling capacity. Storing energy at the temperature of ice requires refrigeration equipment

that can cool the charging fluid (typically, a water/glycol mixture) to temperatures below the

normal operating range of conventional air-conditioning equipment. Special ice-making

equipment or standard chillers modified for low-temperature service are used. When ice thermal

storage is incorporated into a building system the low temperatures of the chilled-water supply

allow the use of low-temperature air distribution (usually calling for Fahrenheit temperatures in

the mid-40s, versus the mid-50s for conventional systems), meaning smaller fans and ducts are

needed. When ice is the storage medium, there are several technologies available for creating ice

and using the ice to cool circulated fluid. The first of which is ice harvesting. Ice harvesting

systems have an evaporator surface on which ice is formed. It is then frequently released into a

storage tank that is partially filled with water. The second is external melt ice-on-coil systems.

They use submerged pipes through which a refrigerant or secondary coolant is circulated. This

causes ice to accumulate on the outside of the pipes. Storage is discharged by circulating the

warm return water over the pipes, melting the ice from the outside. Internal melt ice-on-coil

systems also feature submerged pipes on which ice is formed. Storage is discharged by

circulating warm coolant through the pipes, melting the ice from the inside. Ice slurry systems

store water or water/glycol solutions in a slurry state—a partially frozen mixture of liquid and

ice crystals that looks much like a frozen fruit smoothie. To meet cooling demand, the slurry

may be pumped directly to the load or to a heat exchanger that cools a secondary fluid that

circulates through the building's chilled-water system. Internal melt ice-on-coil systems are the

most commonly used type of ice storage technology in commercial applications. External melt


Sarasota, FL


and ice harvesting systems are more common in industrial applications, although they can also

be applied in commercial buildings and district cooling systems. Ice slurry systems have not

been widely used in commercial applications. Eutectic salts, also known as phase-change

materials, use a combination of inorganic salts, water, and other elements to create a mixture that

freezes at a desired temperature. The material is encapsulated in plastic containers that are

stacked in a storage tank through which water is circulated. The most commonly used mixture

for thermal storage freezes at 47°F, which allows the use of standard chilling equipment to

charge storage, but leads to higher discharge temperatures.

Storage tanks must have the strength to withstand the pressure of the storage medium, and they

must be watertight and corrosion-resistant. Aboveground outdoor tanks must be weather-

resistant. Buried tanks must withstand the weight of their soil covering and any other loads that

might occur above the tank, such as parked cars. Tanks may also be insulated to minimize

external condensation and thermal losses, which typically run 1 to 5 percent per day. Options for

tank materials include steel, concrete, and plastic. Large steel tanks have the capacity of up to

several million gallons. Cylindrical pressurized tanks are generally used to hold between 3,000

and 56,000 gallons. Concrete tanks may be precast or cast in place. Precast tanks are most

economical in sizes of one million gallons or more. Plastic tanks are typically delivered as

prefabricated modular units. Rectangular tanks are commonly available in sizes up to 8 x 8 x 20

feet. Steel and concrete are the most commonly used types of tanks for chilled-water storage.

Most ice harvesting systems use site-built concrete, external-melt systems usually use concrete

or steel tanks, internal melt systems usually use plastic or steel, and eutectic salt systems

commonly use concrete tanks with polyurethane.


Sarasota, FL


Figure 3.2-2 Volume Needed Per Ton-hour

Thermal Storage For This Project:

The Florida State University Ringling Conservation Center is served by a chiller plant that serves

the entire Ringling campus. Currently there are two 650 ton water cooled centrifugal chillers

providing 40 degree water at a 10 degree delta temperature. Using thermal storage on a campus

normally works well due to the fact that they are used year round. Also, when using a campus

instead of just one building the load profiles are normally more stable and a larger on-peak/ off-

peak load exists allowing for enough off-peak cooling to occur to make the option cost effective.

The analysis done on the chiller plant will be to analysis load leveling as well as partial storage

to find which application works best.

Calculation—Partial Storage

All calculations and specifications for equipment talked about in this section can be seen in

Appendix C.

Loads for June, July and August where acquired from the designing engineer. These months

were the only months needed to design the original plant. July, being the design month led to the

design containing one 650 ton chiller with one redundant 650 ton chiller. To do a typical year


Sarasota, FL


analysis bin data was used for each day of the year. Using the relationship between wet bulb

temperatures and tons of cooling a bin year was created.

Once the bin year was created the electricity on-peak and off-peak times (see Figure 3.2-3)

where used to find the maximum storage that would be needed to meet the peak demands while

shifting as much load to off-peak hours as possible. Originally July was assumed as the design

month. Since the off-peak/on-peak times vary between November-March and April-October the

design month actually turned out to be March. This is because March had the largest load where

only 8 hours where available for storage. The other off-peak hours from November to March are

during the day when the ice storage system would be used for meeting the load and not for

storage.

The thermal storage load was determined by subtracting the peak loads from hours 9-19 from the

base load design. The base load is around 110 tons. This chiller was sized at 150 tons. Extra

capacity was added to the system for future expansion.. A York water cooled Centrifugal Chiller

was selected to cover the base load. The thermal storage capacity needed was determined to be

4000 ton hours for the month of March. This load could be stored over eight hours. This

corresponds to a 500 ton thermal storage chiller. A MaxE York Centrifugal Chiller was selected

as the thermal storage chiller. This machine has the capability of making ice, although the

overall efficiency of the chiller is de-rated approximately 30%. The peak ton hours of 4445.5

occur on July 29th. This load could be stored over 11 hours and therefore led to a lower ton per

hour than was needed in March. The storage tanks needed where then sized using Calmac

Icetanks. They were sized off of the max ton hours (4445.5 tonh). It was determined that 8--

1500CSF model tanks are needed. They have the capacity to hold 4560 ton hours of cooling.

Next, the total electrical costs for the system with thermal storage and without thermal storage

were calculated. It was assumed that the 650 ton chillers would run at an ERR (Energy

efficiency ratio = Btu/kW) of 7 when performing the base load and an EER of 13 when

producing the peak load.


Sarasota, FL


July

0.0

100.0

200.0

300.0

400.0

500.0

600.0

700.0

800.0

0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24

Hour of the Day

Tons July

Figure 3.2-3 Peak Day Storage—Partial Storage

Storage Tanks

The thermal storage system chosen was an internal melt ice-on coil system since it is most

commonly used for these applications. Figure 3.2-4 shows pictures taken from the ASHRAE

Thermal Storage Design Guide on how the melt ice-on coil system works.

Base Load Chiller


Sarasota, FL


Figure 3.2-4 Thermal Storage Design Guide—Melt Ice-on Coil System

Electric Rates

Florida power and light will supply the electricity for the chiller plant. A full layout of their

charges can be seen in Appendix C. A summary of the charges used for calculations are listed

below in Table 3.2-1. Using thermal storage led to a kW demand between 500-1999 while

without thermal storage the chiller plant is in the 2000+ kW demand category. The on-peak and

off-peak hours can be seen in Figure 3.2-5.


Sarasota, FL


kW Demand Monthly Demand

Charge On Peak kW

Charge On Peak kWh

Charge Off Peak

kWh Charge 21-499 $38.58 $8.16 3.337¢ 1.021¢

500 -1999 $38.12 $8.15 2.279¢ 0.788¢ Table 3.2-1 Summary of Rate Charges Florida Power and Light

Figure 3.2-3 Summary of On-Peak/Off-Peak Hours

Thermal Storage Conclusions—Partial Storage

From the calculations that can be seen in Appendix C it can be seen that the use of thermal storage

can reduce the cost of running this system by about 52%.

Month

Electric Bill w/out Thermal

Storage

Electric Bill w/ Thermal Storage Savings

January $5,882 $3,291 $2,591 February $6,199 $3,126 $3,073

March $5,913 $3,460 $2,453 April $7,806 $3,845 $3,961 May $7,694 $4,252 $3,441 June $8,331 $4,132 $4,199 July $8,624 $4,422 $4,202

August $8,513 $4,254 $4,259 September $8,308 $4,162 $4,146

October $8,149 $4,194 $3,955 November $6,477 $3,598 $2,879 December $6,186 $3,311 $2,875

Total Price $88,083 $46,047 $42,035

Table 3.2-2 Totals Partial Storage Savings


Sarasota, FL


Table 3.2-2 above shows the overall savings totals per month that can be acquired using thermal

storage. The average savings from November to March is $2774 and from April to October is

$4023. The reason there are two averages given is that during the months of November and March

there are different on-peak off-peak structures for these months.

Month

Without Thermal Storage

Thermal Storage Difference

January $3,799.53 $892.70 $2,906.83 February $4,263.00 $891.61 $3,371.39

March $3,697.04 $891.61 $2,805.43 April $4,466.00 $891.61 $3,574.39 May $4,709.60 $891.61 $3,817.99 June $4,709.60 $891.61 $3,817.99 July $4,679.56 $891.61 $3,787.95

August $4,709.07 $892.70 $3,816.37 September $4,583.74 $891.61 $3,692.13

October $4,486.30 $891.61 $3,594.69 November $4,141.20 $891.61 $3,249.59 December $4,100.60 $891.61 $3,208.99

Total $52,345.23 $10,701.51

% of Overall Price 59.4% 23.2% Table 3.2-3 On-Peak kW Demand Charge

As it can be seen in this table the largest charge is the kilowatt demand charge. When the peak

load occurs during peak hours the cost is significantly larger. The peak kilowatts for each month

can be seen in Table 3.2-4 below. Due to the fact the peak load occurs during on-peak hours for

the chiller without thermal storage the demand charge per month is much greater.


Sarasota, FL


Month Without Thermal

Storage Thermal Storage Difference

January 466 109 357 February 525 109 416

March 455 109 346 April 550 109 441 May 580 109 471 June 580 109 471 July 576 109 467

August 578 109 468 September 565 109 455

October 553 109 443 November 510 109 401 December 505 109 396

Table 3.2-4 On-Peak kW Demand

The peak load now is from the base load chiller and is drastically reduced. Lowering this

kilowatt demand now places the plant in a slightly higher customer service charge bracket.

Shifting the load also puts the plant in a higher kilowatt hour rate for both off-peak and on-peak.

Despite this the overall cost of the thermal storage chiller plant energy usage is still less. As it

can be seen in Table 3.2-5 this shift in demand can be seen in the kilowatt hour charge too.



January $1,081.85 $665.09 $416.76 February $1,001.86 $602.39 $399.47

March $1,153.22 $683.90 $469.32 April $2,625.91 $889.90 $1,736.01 May $2,945.89 $992.32 $1,953.58 June $2,716.93 $903.92 $1,813.01 July $3,089.13 $997.50 $2,091.63

August $2,842.90 $971.13 $1,871.76 September $2,924.83 $898.54 $2,026.30

October $2,879.47 $989.99 $1,889.48 November $1,211.23 $702.43 $508.80 December $1,080.23 $662.79 $417.44

Total $25,553.45 $9,959.90

% of Overall Price 29.0% 21.6% Table 3.2-5 On-Peak kWh Charge


Sarasota, FL


The off-peak charges are higher for the case using thermal storage. This is because the majority

of the load was shifted to off-peak, under the higher demand charge savings was still achieved.

Month



January $962.84 $1,695.02 -$732.18 February $896.51 $1,593.68 -$697.17

March $1,024.83 $1,845.76 -$820.94 April $676.25 $2,024.91 -$1,348.66 May $764.14 $2,329.71 -$1,565.57 June $866.08 $2,298.01 -$1,431.92 July $817.25 $2,494.37 -$1,677.11

August $923.36 $2,842.90 -$1,919.54 September $760.90 $2,333.03 -$1,572.14

October $744.93 $2,273.47 -$1,528.54 November $1,086.26 $1,965.25 -$878.98 December $967.38 $1,718.37 -$750.99

Total $10,490.74 $25,414.47

% of Overall Price 11.9% 55.2% Table 3.2- 6 Off-Peak kWh Charge

Summary Tables 3.2-7, 8 & 9 show the overall shifting of the total kilowatt hours for the entire

year.

Month



January 67,933 157,733 -89,800 February 59,078 140,618 -81,540

March 69,081 165,610 -96,529 April 81,270 198,326 -117,057 May 91,816 228,179 -136,363 June 104,722 225,074 -120,352 July 95,059 236,146 -141,087

August 111,688 237,983 -126,294 September 91,434 228,505 -137,070

October 89,534 222,671 -133,137 November 73,759 178,531 -104,772 December 67,654 159,948 -92,294

Totals 1,003,029 2,379,325 -1,376,296 Table 3.2-7 Off-Peak kwh


Sarasota, FL




January 97,864 28,213 69,651 February 87,491 24,472 63,018

March 107,710 35,664 72,046 April 115,222 28,559 86,663 May 129,263 29,737 99,526 June 119,216 29,026 90,190 July 138,935 38,052 100,884

August 124,743 30,072 94,671 September 128,338 28,855 99,483

October 126,348 29,667 96,681 November 106,409 27,496 78,913 December 98,657 28,216 70,440

Totals 1,380,196 358,030 1,022,166 Table 3.2-8 On-Peak kwh

Category Without Thermal

Storage Thermal Storage

Total kWh 2,383,225 2,737,355 kWh On-Peak 1,380,196 358,030 kWh Off-Peak 1,003,029 2,379,325

% kWh On-Peak 57.9% 13.1% % kWh Off-Peak 42.1% 86.9%

Table 3.2-9 Summary of kwh Usage

The thermal storage chiller will be laid out with the chiller downstream as shown in Figure 2-7.

Using the chiller downstream allows the warm water to first flow through the storage tanks

cooling it down before it enters the chiller. This arrangement results in a higher usable storage

capacity and a constant discharge temperature. The downside to using this arrangement is that

when the chiller operates at a lower temperature it is less efficient.


Sarasota, FL


Figure 3.2-7 Chiller Configuration

Payback Period on Investment

The payback period for the partial thermal storage system was calculated. The first cost of the

original system and the thermal storage system are shown in Tables 3.2-10, 11 below.

Equipment Thermal StorageIce Chiller 500 tons $250,000.00 Base Load Chiller 180 ton $94,250.00 Redundant Chiller 500 Tons $169,600.00 Ice Storage Tanks (4560 Ton-hr) $456,000.00

Total $969,850.00 Table 3.2-10 Thermal Storage First Cost

Equipment Original System 650 Ton Chiller $215,000.00 651 Ton Chiller Redundant $215,000.00

Total $430,000.00 Table 3.2-11 Original System First Cost

The difference in first cost between these two systems is $754,850.00. Using the inflation rate of

money to be 1.023% per year a payback period of 12.01 years was calculated.


Sarasota, FL


Calculation—Load Leveling

All calculations and specifications for equipment talked about in this section can be seen in

Appendix C.

The calculations for the load leveling system were done using the same bin year created in the

partial storage section. These loads were then compared the original chiller plant design as done

for the partial storage system. The design day for this system occurred in July. The thermal

storage capacity needed was determined to be 2650 ton hours for the design day in July. This

corresponds to a 310 ton thermal storage chiller. A MaxE York Centrifugal Chiller was selected

as the thermal storage chiller. The storage tanks needed where then sized using Calmac

Icetanks. It was determined that 7-- 1320CSF model tanks are needed. They have the capacity to

hold 2660 ton hours of cooling. Figure 3.2-8 shows the design day load.

July

0.0

100.0

200.0

300.0

400.0

500.0

600.0

700.0

800.0

0 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19 20 21 22 23 24

Hour of the Day

Tons July

Figure 3.2-8 Design Day Storage—Load Leveling

STORE


Sarasota, FL


Thermal Storage Conclusions—Load Leveling

From the calculations that can be seen in Appendix C it can be seen that the use of thermal

storage can reduce the cost of running this system by about 15.8%. The overall cost analysis can

be seen in the following tables. The results are similar to the partial storage system but less

savings is achieved because using load leveling shifts less kilowatt usage to off-peak than the

partial storage system did. Again the thermal storage chiller will be laid out with the chiller

downstream as shown in Figure 3.2-7.

Month

Electric Bill w/out Thermal

Storage

Electric Bill w/ Thermal Storage Savings

January $5,821 $5,262 $559 February $6,619 $6,012 $607

March $5,707 $5,185 $522 April $7,529 $5,985 $1,544 May $8,005 $6,426 $1,578 June $8,111 $6,632 $1,479 July $8,199 $6,864 $1,335

August $8,115 $6,412 $1,703 September $7,960 $6,302 $1,658

October $7,854 $6,287 $1,567 November $6,450 $5,889 $561 December $6,317 $5,693 $624

Total Price $86,687 $72,949 $13,738

Table 3.2-12 Totals Load Leveling Savings

Table 3.2-12 above shows the overall savings totals per month that can be acquired using load

leveling thermal storage. The average savings from November to March is $575 and from April

to October is $1552. Again the reason there are two averages given is the two rate structures that

occur.


Sarasota, FL


Month



January $3,799.53 $2,170.96 $1,628.57 February $4,278.75 $2,480.26 $1,798.49

March $3,710.70 $2,138.95 $1,571.74 April $4,482.50 $2,138.95 $2,343.55 May $4,727.00 $2,570.08 $2,156.92 June $4,727.00 $2,652.31 $2,074.69 July $4,696.85 $2,744.90 $1,951.94

August $4,709.07 $2,564.26 $2,144.81 September $4,600.68 $2,520.51 $2,080.17

October $4,502.88 $2,514.42 $1,988.45 November $4,156.50 $2,429.68 $1,726.82 December $4,115.75 $2,348.57 $1,767.18

Total $52,507.19 $29,273.85

% of Overall Price 60.6% 40.1% Table 3.2-13 On-Peak kW Demand Charge Load Leveling

As in the partial storage case the largest charge is the kilowatt demand charge. The peak

kilowatts for each month can be seen in Table 3.2-14 below. As you can see, less kilowatts are

able to be shifted using load leveling as opposed to the partial storage case.



January 466 266 200 February 525 304 221

March 455 262 193 April 550 293 257 May 580 315 265 June 580 325 255 July 576 336 240

August 578 314 264 September 565 309 256

October 553 308 244 November 510 298 212 December 505 288 217

Table 3.2-14 On-Peak kW Demand Load Leveling

Lowering this kilowatt demand again places the plant in a slightly higher customer service

charge bracket. Shifting the load also puts the plant in a higher kilowatt hour rate for both off-


Sarasota, FL


peak and on-peak. As it can be seen in Table 3.2-15 this shift in demand can be seen in the

kilowatt hour charge too.

Month



January $1,065.88 $1,917.66 -$851.78 February $1,251.60 $2,190.87 -$939.27

March $1,059.21 $1,889.38 -$830.17 April $2,412.87 $2,378.69 $34.18 May $2,594.39 $2,553.99 $40.40 June $2,688.41 $2,635.70 $52.70 July $2,780.95 $2,727.72 $53.23

August $2,689.47 $2,548.21 $141.26 September $2,662.58 $2,504.73 $157.85

October $2,649.98 $2,498.68 $151.30 November $1,229.35 $2,146.19 -$916.84 December $1,166.39 $2,074.55 -$908.16

Total $24,251.07 $28,066.37

% of Overall Price 28.0% 38.5% Table 3.2-15 On-Peak kWh Charge Load Leveling

The off-peak charges are higher for the case using thermal storage. This is because the majority

of the load was shifted to off-peak. Under the higher demand charge savings was still achieved

though.


Sarasota, FL


Month



January $956.09 $1,173.47 -$217.38 February $1,088.62 $1,340.65 -$252.03

March $936.68 $1,156.17 -$219.48 April $633.53 $1,212.99 -$579.45 May $683.39 $1,302.38 -$618.99 June $695.71 $1,344.05 -$648.34 July $720.80 $1,390.97 -$670.17

August $716.58 $1,299.43 -$582.85 September $696.95 $1,277.26 -$580.31

October $701.21 $1,274.18 -$572.97 November $1,064.52 $1,313.31 -$248.80 December $1,034.78 $1,269.47 -$234.69

Total $9,928.84 $15,354.32

% of Overall Price 11.5% 21.0% Table 3.2- 16 Off-Peak kWh Charge Load Leveling

Summary Tables 3.2-17, 18, 19 show the overall shifting of the total kilowatt hours for the entire

year.

Month



January 97,864 74,228 23,636 February 87,491 76,596 10,895

March 107,710 73,133 34,577 April 115,222 79,203 36,019 May 129,263 87,874 41,388 June 119,216 87,760 31,456 July 138,935 82,104 56,831

August 124,743 87,675 37,068 September 128,338 83,399 44,939

October 126,348 85,971 40,377 November 106,409 80,394 26,015 December 98,657 80,300 18,356

Totals 1,380,196 978,638 401,558 Table 3.2-17 On-Peak kwh Load Leveling


Sarasota, FL


Month



January 67,933 123,713 -55,780 February 59,078 127,660 -68,582

March 69,081 121,889 -52,807 April 81,270 132,004 -50,735 May 91,816 146,457 -54,641 June 104,722 146,267 -41,545 July 95,059 156,419 -61,360

August 111,688 146,125 -34,437 September 91,434 143,632 -52,198

October 89,534 138,663 -49,129 November 73,759 133,990 -60,231 December 67,654 133,834 -66,180

Totals 1,003,029 1,650,653 -647,624 Table 3.2-18 Off-Peak kwh Load Leveling

Category


Thermal Storage

Total kWh 2,383,225 2,629,290 kWh On-Peak 1,380,196 978,638 kWh Off-Peak 1,003,029 1,650,653

% kWh On-Peak 57.9% 37.2% % kWh Off-Peak 42.1% 62.8%

Table 3.2-19 Summary of kwh Usage Load Leveling

Payback Period on Investment

The payback period for the load leveling thermal storage system was calculated. The first cost

of the original system and the thermal storage system are shown in Tables 3.2-20 & 21 below.

Equipment Thermal Storage Ice Chiller 320 tons $160,000.00 Redundant Chiller 650 Tons $215,000.00 Ice Storage Tanks (2660 Ton-hr) $266,000.00

Total $641,000.00 Table 3.2-20 Thermal Storage First Cost Load Leveling

Equipment Original System 650 Ton Chiller $215,000.00 650 Ton Chiller Redundant $215,000.00

Total $430,000.00 Table 3.2-21 Original System First Cost Load Leveling


Sarasota, FL


The difference in first cost between these two systems is $211,000.00. Using the inflation rate of

money to be 1.023% per year a payback period of 15.52 years was calculated.

Thermal Storage—Comparing Systems

Both the partial storage and load leveling systems have too great of payback periods to make

them worth investing in. The systems both have similarities and differences in results. First,

both systems successfully shifted kilowatt hours from on-peak to off-peak hours. However the

partial storage system, which is really set up as a base load chiller and then full storage of the

rest of the load, is more successful at shifting kilowatt hours. With using load leveling about 235

kW of on-peak demand were able to be shifted to off-peak as well as 53,970 kWh on a monthly

average. With using partial storage about 430 kW of on-peak demand were able to be shifted to

off-peak as well as 114,690 kWh on a monthly average. Therefore the savings for the partial

storage system is much greater. Secondly, the shift for both systems is enough to move the plant

down into the next pay structures with Florida Power in Light. This actually really hurts the cost

savings of both systems. The on-peak kwh charge in the lower bracket is 1.058 cents higher than

the higher bracket. The off-peak kwh charge in the lower bracket is 0.233 cents higher. The on-

peak kW demand charge only changes slightly from $8.15/kW to $8.16/kW. Had the plant

stayed in the same bracket the savings would have been much greater for both of these systems.

Finally, the first cost of a thermal storage system in both cases is too high for the amount of

money the plant is able to save on energy bills. The main extra charge comes from the storage

tanks at $100/tonh. Using load leveling less ton hours are needed and therefore the cost of the

tanks is less. However storing less ton hours results in a higher number of on-peak hours which

in turn leads to less savings for the overall system. Finally, both systems overall use more

kilowatts with thermal storage than they do without thermal storage. This does not meet the

design objectives to reduce energy usage. Thermal storage reduces costs by shifting time of use.

The decrease in energy efficiency in making ice increases the overall energy consumption of the

plan.


Sarasota, FL


Chiller Configuration

Currently the chiller plant contains two 650 ton chillers. One of these chillers is redundant.

There are plans to add possibly one or two more buildings onto the plant in the future. The

Ringling campus is running out of land though. Looking at the overall load that the chillers

currently see during about 85% of the year the tonnage needed is under 500 tons. Over night the

load reduces to around 110 tons. A 650 ton chiller running at 110 ton has a very poor efficiency.

If the system had 2—350 ton chillers and 1—300 ton chiller it would be about to meet the loads

with a better efficiency. Also this set up will allow the redundant chiller to be 350 tons therefore

saving on first cost. If more buildings are added to the system more than likely the 300 ton

chiller would still be good for the over night base load. In both set ups another chiller will

probably have to be added to the system when more buildings are added. With this new

configuration the chillers will run closer to full load more often allowing for the overall

efficiency of the system to be higher.

First Cost and Operation Cost Comparison

Equipment First Cost 650 Ton Chiller $215,000.00 650 Ton Redundant $215,000.00

Total $430,000.00 Table 3.2-22 Original Design First Cost

Equipment First Cost 300 Ton Chiller $89,900.00 350 Ton Chillers $121,800.00 350 Ton Redundant Chiller $121,800.00

Total $333,500.00 Table 3.2-23 New Chiller Configuration

From this first cost analysis it can be seen that just by using three smaller chillers opposed to two

larger chillers reduces the first cost of the equipment by $96,500. On a first cost analysis this

makes this option more attractive. Next, the operation cost of running these two configurations

was analyzed. As before it was assumed that on average the 650 ton chiller would run at an EER

of 7 during the night when the load is around 110 tons and at on average an EER of 13 during

the day. With the new configuration the chillers would run on average at an EER of 13 all the


Sarasota, FL


time. This led to the cost savings of approximately $8,040 a year in operating costs. An overall

cost of operation of each configuration and the savings achieved by the new configuration can be

seen in Table 3.2-24 below. On average from November to March the new system saves $524 a

month and from April to October $775 per month.

Table 3.2-24 Operation Cost Saving New Chiller Configuration

Month Cost with 2-650

Ton Chillers

Cost with 1-300 Ton, 2-350

Ton Chillers Savings January $5,882 $5,492 $390 February $6,199 $5,662 $537

March $5,913 $5,433 $480 April $7,806 $6,993 $814 May $7,694 $7,289 $405 June $8,331 $7,451 $879 July $8,624 $7,828 $797

August $8,513 $7,644 $869 September $8,308 $7,461 $847

October $8,149 $7,339 $810 November $6,477 $5,694 $782 December $6,186 $5,757 $430

Total Price $88,083 $80,043 $8,040

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