Department of Emergency and Military Affairs
United States Army
Constructed in 1993, the facility is replete with energy conservation
strategies of the time. Variable pumping and airflow schemes,
variable flow capabilities of cooling tower fans, and four-pipe
hot water and chilled water distribution systems should have
made for an efficient HVAC system. Direct Digital Controls (DDC)
have since replaced outdated controls, and should have been
effective in generating even more utility cost reductions.
as is often the case, the facility’s HVAC system was substantially
oversized and as a result required exacting controls capabilities
for efficient operation, yet a lack of commissioning and improper
application of controls strategies exacerbated the problems.
In most cases, a “bigger is better” approach to HVAC systems
is counterproductive and results in the system’s using excessive
energy and creates poor space conditions. Oversized HVAC systems
typically have any number of deficiencies, but the good news
is that once solutions have been devised and implemented the
energy costs fall dramatically and space conditions improve
of historical energy data is rather disturbing. Our HVAC analysis
did not account for the 100-HP hydraulic pump set serving the
flight simulators, yet even so the energy consumption is staggering
if compared to like-use facilities with similar HVAC systems.
Utility charges include .0615$/KwH and $15.25/Kw for demand.
The demand for the facility has ranged from 534 to 403 KW over
the past 3 years, a very high number, but not surprising considering
how many HVAC electrical loads – fan motors, pumps, water chillers,
cooling tower fans - are operating unnecessarily to condition
building occupancy profile is essentially office with air-conditioned
“warehouse” space (the simulator bays) and their supporting
computer rooms. As will be discussed later, the electrical loads
(and resultant HVAC loads) in the computer rooms are quite small.
Operating schedules for the hydraulic pump set and the ability
for continuous operations (8,760 hours per year) make estimates
an inexact science. Nevertheless, based upon our experiences,
considering the energy conserving devices used by the HVAC system
and the loads they serve, and making assumptions for the hydraulic
pump set, we estimated that the facility was needlessly expending
a significant amount of its current utilities.
Subject to OptimissioningSM
Generally speaking, the required work for this project focused
exclusively on the chilled water aspects of the heating, ventilating
and air conditioning (HVAC) system, with emphasis on the operation
of the computer room a/c (CRAC) units relative to both mechanical
refrigeration and waterside economizer modes of air conditioning.
It was expected that the determination of staging for the CRAC
units would be an iterative process to determine the most efficient
method of control: staging on multiple units at higher supply
air temperatures, running the minimum number of units at lower
supply air temperatures, or a combination of the modes as is
variation in supply air temperatures from the CRAC units will
be a function of the chilled water supply temperature. Currently,
warmer than design water is continuously delivered to the units,
caused by poor piping practices which are creating a reverse
decoupling of the secondary chilled water. As now intended,
the chilled water temperature will be warmer intentionally,
either due to the use of a chilled water reset schedule for
the water chillers, or operation of the plate frame heat exchanger.
Warmer water will entail warmer supply air, necessitating additional
air, which is to be supplied by the redundant CRAC units staged
by soon to be installed controllers.
In addition to the computer rooms CRAC units, other systems
• The entire chilled water delivery system, including the three
water chillers, three primary pumps, and two secondary pumps
that are equipped with variable frequency drives.
• The condenser water system including the three condenser water
pumps and three cooling towers and their respective fans equipped
with variable frequency drives.
• The air distribution system, particularly VAV air handling
unit #3 and all associated VAV boxes.
• Air handling units #1, 2, 4 and all fan coil units will be
evaluated for cooling load requirements and chilled water and
air flow rates will be adjusted during the rebalance phase of
• The entire DDC system will be evaluated and modified if deemed
Chilled water and condenser water system:
• Three (3) nominal 110-ton reciprocating water-cooled water
chillers selected for 55-degree entering, 45-degree leaving
• Three (3) cooling towers selected for 95-degree entering,
85-degree leaving condenser water temperatures with 340 GPM
at 79-degree wet bulb (WB) temperature. Each tower fan is equipped
with a variable frequency drive.
• Three (3) 5 HP chilled water pumps selected for 270 GPM at
40’ total head pressure.
• Two (2) 20 HP secondary chilled water pumps equipped with
variable frequency drives, selected for 810 GPM and 50’ of head
each. One pump is redundant.
• Three (3) 5 HP condenser water pumps selected for 340 GPM
at 40’ total head pressure.
• One (1) plate/frame heat exchanger, approximately 73 tons
capacity. Heat exchanger is selected for a 3-degree approach,
with 47 degrees entering the heat exchanger and 50 degrees leaving
to the chilled water system.
There are four (4) air-handling units providing space temperature
control. Air handling unit #1 serves the East bay of the facility,
air handling units #2 and #4 serve the West simulator bay(s).
This bay (West) is only now being outfitted with simulators,
and the corresponding computer room put into service – including
the installation of all computing hardware - to provide the
requisite computer simulations for a fully functional simulator
air-handling units’ chilled water-cooling coils were selected
(design day) for 45-degree entering and 55-degree leaving water
temperature, with a resultant deck temperature of approximately
55-58 degrees. Air handling units #1, #2, and #4 are equipped
with hot water coils, and are constant airflow. Air handling
unit #3 serves the second floor office areas and is variable
flow, utilizing the aforementioned VAV boxes for zone temperature
control. Four-pipe fan coil units serve the remaining zones’
(essentially the “Lobby” areas) heating and cooling requirements.
The dual “computer rooms” are an interesting and divergent subset
of the HVAC operational challenges at the facility. Both rooms
appear to have at least one redundant computer room air conditioning
unit (CRAC) installed. The East room currently uses four (4)
nominal 22-ton units, and the West room utilizes six (6) nominal
15-ton units and has two (2) additional air-cooled (remote condenser)
DX units, nominal 5 tons each.
rooms utilize and under-floor supply air delivery system, and
all CRAC units are complete with chilled water and hot water
coils, infra-red (IR) humidifiers, and have relatively simple
temperature and humidity controls. Both the chilled water and
hot water control valves are 2-way modulating.
The current control system is Direct Digital Control (DDC),
manufactured by Alerton Controls, and is currently being maintained
by Climatec, who also installed and commissioned the system.
Our observations of the control system indicate that it is quite
capable of providing the level of control required to maintain
comfortable space conditions, adequate chilled water and hot
water capacity for computer room operations, and do so efficiently.
system is a fairly sophisticated computerized distributed processing
system that uses “Visual Control Graphical Programming” language.
This language is extremely versatile when an experienced programmer
designs the functional programs, provided the correct sensors
and inputs are in place. The resolution of the temperature/analog
(input) sensor(s) appear to be more than sufficient to provide
excellent temperature control and system response
Deficiencies and Solutions
The system had a 43-degree chilled water setpoint, considerably
below the design leaving of 45 degrees leaving, and approximately
48-degree return chilled water – an obvious indicator that the
system is severely over-pumped. Moreover, it appeared as if
approximately 46-degree water is actually being delivered to
the chilled water coils throughout the facility, almost certainly
due to “reverse” decoupling thru the primary-secondary decoupler.
In other words, the return chilled water – itself extremely
low at approximately 48 degrees – was mixing with the 41 degree
chilled water from the chillers, affecting a delivered chilled
water temperature of approximately 46 degrees.
reverse decoupling is common for primary-secondary piped systems
that employ full-size, front decoupled methodology. It was evident
that one of the two available secondary pumps is intended to
be redundant, yet both were operating and both are operating
at very high speeds to maintain the differential pressure setpoint.
This high DP setpoint was causing other energy wasting, space
temperature control issues including affecting the chilled water
control valves’ ability to modulate flow effectively. The location
of the differential pressure (DP) sensor was also observed to
be less than ideal, and is controlling to much too high of a
high chilled water system pressures were causing such issues,
and the resulting performance penalty was three-fold. First,
sub-cooling of the air will require additional reheat to maintain
zone temperature setpoint, obviously an inefficient operation.
Second, some zones may not have the ability to achieve setpoint
during the winter (heating) seasons, creating uncomfortable
conditions and likely occupant complaints. Third, the pumps
must run at higher speeds, using more energy, to maintain the
high DP setpoint. Conversely, in summer (cooling) months the
inability to deliver cold enough water, due to the reverse de-coupling
described above, in order to dehumidify the air will result
in “stuffy” space conditions. Comfort and productivity issues
have been well documented to be directly proportional, so the
net result becomes higher energy costs and lower productivity.
Water System Deficiencies:
The initial study of the chilled water system revealed that
the entire system was grossly oversized, poorly designed and
completely out of control. Our estimates indicated that approximately
120 tons of chiller capacity would be required on a design day,
yet 330 tons was actually installed. As discussed above, an
oversized system is typically counterproductive from an energy
perspective unless a concentrated effort is made to eliminate
wasteful operating practices.
visits revealed that the chillers had different setpoints: Chiller
1B was operating with a setpoint of 43- degrees, had 12,715
starts and 25,313 total run hours. Chiller 1A’s setpoint was
41-degrees, had 25,639 starts and 29,459 run hours. Chiller
1C also had a 41-degree setpoint, 23,630 starts and 29,452 total
run hours (all above data thru 2003). The number of starts per
run hour is cause for concern, as the units are obviously starting/stopping
much too frequently, and the total run-hours for all (3) machines
shows that 330 tons is commonly being produced to satisfy a
120 ton load. This phenomenon would be expected from a system
that experiences very low chilled water system temperature differentials
such as existed at this facility.
the original plate/frame heat exchanger was selected for a 3-degree
approach between entering cold (condenser water) and leaving
hot (chilled water supply), limiting the number of hours that
the facility could be satisfied using “free” cooling. It was
also determined that the existing control sequence of operations
had several flaws which would prohibit plate/frame heat exchanger
operations regardless if the ambient conditions were suitable
for “free” cooling operation.
Water System Solutions:
Installing a new 2-position isolation valve in the existing
de-coupler, and installing a new, smaller sized de-coupler in
a much better location – from the end of the primary pumps’
header to the main return line - solved the “reverse” de-coupling
issues. A bi-directional flow meter was installed in the de-coupler
to monitor both flow rates and direction of flow. As expected
and intended, the flow thru the de-coupler is now from supply
to return, depressing the return water temperature and allowing
the chillers to unload.
new 1-degree approach plate/frame heat exchanger was installed
to replace the existing 3-degree heat exchanger, extending potential
operating hours. During the optimissioningSM phase, the chilled
water diverting valve responsible for directing the chilled
water flow to the heat exchanger was found to be defective and
was subsequently replaced.
central plant sequence of controls was rewritten to correct
existing deficiencies, and applicable setpoints and timeouts
– i.e. the time delays going into and coming out of plate/frame
and chiller operation - were modified and optimissionedSM. As
the HVAC system is substantially oversized, duty cycles were
established for all chillers, cooling towers, and primary and
secondary pumps as shown below. A chilled water temperature
reset schedule was established based the ambient wet-bulb (WB)
heat exchanger setpoint and the chilled water reset schedule
upper limit are close in proximity to the actual deliverable
chilled water supply temperature. The cooling towers have approximately
a 6-degree approach, and coupled with the 1-degree heat exchanger
installed, the chilled water supply temperature at 52 degree
WB (ambient) will result in approximately 59-degree supply chilled
water. Above 52-degree WB, the chillers will start and produce
approximately 57-degree supply chilled water. OptimissioningSM
observations indicate that the Data Rooms, when enabled with
temperature staging discussed below, can maintain temperature
limits at these supply chilled water temperatures.
chilled water system differential pressure (DP) control was
also optimissionedSM. A system pressure of 12.5 PSIG was being
maintained prior to optimissioningSM the HVAC system. Subsequent
to improvements made in the air systems and CRAC’s as discussed
below, the final system setpoint is now 3.75 PSIG.
to the HVAC improvements and subsequent optimissioningSM, the
chilled water system DT was extremely low (approximately 3-4
degrees) and caused additional chiller systems to be started
in response. This is a common requirement of most systems, as
most experience low system DT’s and the resultant inability
to load the chiller to its full load capacity. For example,
the chillers at CMSB are rated for approximately 113 tons at
a 10-degree (55-45) chilled water DT.
new central plant piping design and new sequence of operations
resulted in the chilled water loop receiving warmer than design
(45-degrees) entering water temperatures during plate-frame
operation. Improving chiller plant - specifically chilled water
piping/pumping/controls - operations eliminated the need for
the Sears dehumidifiers, as the ability to deliver cold enough
chilled water (during high temperature/humidity ambient conditions)
to the CRAC units to dehumidify the computer rooms was greatly
improved. At project closeout, unusually high ambient conditions
existed, yet the entire facility was operating on the plate/frame
heat exchanger for the majority of operating hours. Beginning
on 3/15/2005, the run-hours of the water chillers were reset
to zero so that chiller run times could be tracked from the
completion of the project.
Water System Deficiencies:
During optimissioningSM, it became apparent that the cooling
towers were seriously underperforming. Observations and readings
• Cooling tower fan motors operating at approximately 40% of
rated full load amps even though the variable frequency drives
were at 60Hz.
• One operating cooling tower could not achieve predicted wet-bulb
(WB) approach results of approximately 8-degrees between outside
air WB and supply condenser water temperature. Two cooling towers
were required, ½ of the total condenser water flow each, to
achieve this approach.
• Extremely restricted airflow thru the cooling tower fill,
as would be expected considering the above.
DDC sequences controlling the condenser water system were less
than ideal and essentially precluded the use of the heat exchanger
due to their programming logic and setpoints. Also, during the
optimissioningSM phase, it was discovered that three modulating
condenser water isolation valves were defective; one was located
at cooling tower #1B, one at cooling tower #1C, and one at Chiller
Water System Solutions:
The condenser water valves were replaced, and the cooling towers
were also replaced with an identical model in order to facilitate
an expedient change-out and keep the central plant in operation
at all times. The new cooling towers were incorporated into
the new central plant sequences and at project closeout were
performing extremely well.
cooling towers can produce approximately 6-degree approach to
ambient WB as described above, enabling a quite high heat exchanger
changeover setpoint of 52 degrees. Programming code was added
so that when heat exchanger mode is initiated, two (2) cooling
towers are energized to suppress the cooling tower supply temperature
more quickly than one tower. The cooling tower variable frequency
drives’ PID control loops were modified to eliminate “hunting”,
and the condensing water system is performing as designed and
intended at this time.
timing of critical actions was established during optimissioningSM.
• A 10-minute period was established to allow the condenser
water system to achieve the heat exchanger setpoint
• The 10-minute delay is observed before switching any valves
needed to enable the heat exchanger mode of operation
• Two cooling towers are enabled on a call for heat exchanger,
enabling the condenser water loop to be sufficiently depressed
during the 10-minute delay
new sequence of operations for the DDC system also included
changes to the condenser water system. The cooling towers have
a common supply water header, so any tower can serve and chiller
system. As we expect two cooling tower to be required at the
very most, and one cooling tower required the vast majority
of the time, the cooling tower were set up to rotate Lead-Lag-Standby
on a weekly basis in an attempt to maintain a consistent water
condenser water pumps are dedicated to their respective chillers
and will rotate in conjunction with the chiller duty cycles.
Chiller #1A is to be permanently dedicated as Emergency (Standby),
yet its condenser water pump (#3A) is dedicated to the heat
exchanger and as such is enabled on a call for heat exchanger
Rooms Systems Deficiencies:
Similar to the chilled water system, the Computer Rooms Systems
are also greatly oversized. All of the four nominal 22-ton units
(88 total tons) were in operation in the East Computer Room
during all site visits and prior to any HVAC system improvements
or optimissioningSM, indicating a lack of control capability
and a ready source for energy conservation. Also, a number of
residential style Sears brand dehumidifiers were located throughout
the East room, with personnel indicating that they are required
during the monsoon (high humidity) season. We believe this to
be caused by the “reverse” decoupling discussed earlier, as
the entering chilled water temperature to the CRAC units’ chilled
water coils is warmer than design, and not sufficiently cold
to dehumidify the air.
from the original severs (East room) shows a total heat of rejection
of 487,100 BtuH, or approximately 40 tons. This is assumed to
be the maximum possible heat rejected by all servers, but the
service entrance section (main power) dedicated to the servers
showed many unused circuits. Field measurements indicate an
approximate heat load of 170,750 BtuH, or 14.2 tons in the East
Computer Room, and 102,450 BtuH or 8.5 tons in the West Computer
Room currently under construction. Additional cooling loads
will include lighting, people and infiltration loads, but we
expect that these are minimal. Field measurements of the service
entrances providing power to the server racks is as follows:
A B C Average
E Room: Service 1: Current 95 102 113 103 Amps
Voltage 198 198 198 198 Volts
Service 2: Current 22 28 19 23 Amps
Voltage 198 198 198 198 Volts
Room: Service 1: Current 88 78 62 76 Amps
Voltage 204 204 204 204 Volts
Service 2: Current 10 6 2.4 6 Amps
Voltage 204 204 204 204 Volts
E Room: (103 + 23) * 198 * 1.732 = 43210 VA = 43.2 kW = 13 tons
W Room: (76 + 6) * 204 * 1.732 = 28973 VA = 29.0 kW = 9 tons
East room obviously has cooling capacity to spare – as does
the West room where there are six (6) 7.5-ton units installed
and especially considering the more modern servers recently
installed – but both lack the controls hardware and software
capability required to provide reliable and energy efficient
operations of the CRAC units. The lack of controls capability
for the CRAC units has been and is a root cause of poor performance
and high-energy consumption.
all units operated 24-hours/day, even though two (at least)
in each Data Room are redundant. Unnecessarily operating all
units was obviously wasting energy, but was also generating
a fan energy cooling load of approximately 7,000 - 20,000 Btu/H
per operating unit depending upon the fan motor brake horsepower.
In other words, running all of the CRAC units was actually adding
load to the Data Rooms due to the small actual cooling requirement
and constant volume, in air stream fan/motor combinations.
CRAC units were originally purchased with non-spring return
chilled water valve actuators, and the on-board Liebert controls
are programmed such that the chilled water valve actuators stay
in their respective last position when the unit is de-energized.
Because the chilled water system is primary-secondary pumped,
and is variable flow secondary (using variable frequency drives),
it is highly probable that the chilled water system was flowing
much more chilled water than required – and wasting energy -
when any of the units were staged off.
is also worth noting that the condition of many of the CRAC
units is poor. None of the integral infrared humidifiers were
operable, many units have moderate to severe corrosion (rust),
two of the units had inoperable chilled water valve actuators,
and one had a faulty control board. The plenum space under floor
is also quite dirty, and could possibly lead to premature failure
of the computer hardware as the server racks are supplied with
conditioned air from below the raised floor. Quotations for
repairing the units, cleaning the units and the under-floor
plenum, and establishing a preventative maintenance contract
with a qualified service organization were provided.
Room System Solutions:
Many iterations of the chilled water coils (Liebert CRAC units)
provided sufficient information to formulate and recommend an
integral central plant operating strategy. Further analysis
of the CRAC’s indicates satisfactory operation of the server
racks with approximately 63-degree supply air, which is achievable
with approximately 60-degree entering water temperature (EWT)
and design (52 GPM) flow rates As the supply air temperature
from the CRAC units is paramount to reliable server and resultant
simulator operations, the performance of these chilled water
coils played a vital role in overall system operations, performance
and energy efficiency.
establish control of the CRAC units, a Liebert RAC-8 controller
and temperature sensor were installed in each Data Room. The
RCA-8’s enable the units to be duty cycled, manually over-ridden
and started/stopped depending on space temperature and user
requirements. Energy-Environment-Economics was also able
to develop and implement a method for driving the chilled water
valve actuators closed when the unit is off, enabling a lower
DP setpoint for the secondary pump variable frequency drives.
CRAC units’ modeling during the optimissioningSM phase established
setpoints for staging the CRAC units’ on/off, based upon the
room temperature, that will enable the CRAC the ability to maintain
ample cooling capacity during both chiller and heat exchanger
modes of operation. The duty cycle will rotate the lead CRAC
unit on a 2-week basis. See Section for all Tables.
Handling System Deficiencies:
The air handling unit chilled water coil selections reflect
a total chilled water flow requirement of approximately 209
GPM, indicating a design day load of approximately 87 tons was
the engineered load point for the four (4) zones. Our load calculations
indicate a required peak load of approximately 49 tons, which
equates to approximately 118 GPM based upon a 10-degree water
temperature rise and includes the packaged unit loads serving
the inside of the simulators. Design flows to all Lobby fan
coil units totaled approximately 150 GPM, or approximately 63
tons. Our load calculations show a design load of approximately
25 tons, or 60 GPM for a 10-degree rise coil.
there is a considerable amount of excess capacity, both water
and airflow, available to serve the respective zone loads. Field
measurements of the supply air temperatures were taken, with
the checked air-handling units delivering approximately 62-degree
supply air during typical “winter” conditions. This is important
information for the development of new controls strategies to
utilize the plate-frame heat exchanger and reduce overall energy
load profiles for the air handling unit zones indicates more
than sufficient capacity, even considering that the simulators
in the East and West bays were using nominal 4-ton packaged
heat pumps for simulator (inside) cooling, and rejecting the
heat directly into the bay. Considering the heat of compression,
while operating these heat pumps are adding approximately 10
tons of load to the East bay, requiring the air handling unit(s)
and ultimately the chilled water system to reject this heat
load. The West bay, newly outfitted with simulators, also has
two nominal four (4) ton packaged units rejecting loads directly
to the control zone(s).
general, the air handling unit’s original selections including
airflow quantities, coil selections and control valve selections
were less than ideal. These factors have contributed to the
poor performance and high-energy costs of the complete HVAC
system. For example, the air handling units’ inability to modulate
the chilled water flow effectively has helped to create very
low (approximately 4 degrees) system temperature differential
(DT) and increased pump HP. Low system DT’s also has the ultimate
effect of limiting the loading of the affected chiller(s), necessitating
that additional chiller systems to be started to compensate.
air handling unit #3, equipped with a variable frequency drive,
was discovered to have an unnecessarily low supply air temperature
reset setpoint (50-degrees). Coupled with it’s associated VAV
boxes (apparently) never having been commissioned, the reset
schedule – established to set the supply air temperature required
to satisfy the greatest zone demand - was continuously calling
for 50-degree air even during winter months, requiring that
much more unavailable reheat. Nearly all of the VAV boxes were
severely out of calibration, further exacerbating space temperature
deviations from setpoint and subsequent tenant complaints.
Handling System Solutions:
The chilled water valves were replaced with pressure independent
2-way control valves to reduce the over-pumping problem. Experience
with these valves has been excellent, and the infinitely variable
Cv automatically compensates for fluctuations inherent in DP
controlled variable secondary pumping systems. These valves,
coupled with the central plant revisions and the new sequence
of controls, have more than tripled the system temperature differential
(DT). As evidenced by the DP setpoint (previously 12.5 psig,
now at 3.75 psig) and one (two previously) secondary pump operating
at approximately 40Hz (two at approximately 52 Hz previously),
the chilled water system
Digital Control (DDC) System Deficiencies:
The DDC system was a complete disaster. All of the necessary
capabilities were resident, they simply had not been implemented
into a comprehensive energy management strategy. Setpoints were
unrealistic or too high, PID loops were not tuned, sensors were
out of calibration, reset schedules were too aggressive, not
aggressive enough or non existent.
Digital Control (DDC) System Solutions
All existing sequences were reviewed and either rewritten entirely
or modified. New programming was added, all setpoints were optimized,
existing sensors were calibrated or replaced. Essentially, the
entire DDC system was revamped.
After commencing in November of 2004 and being released to full
control March of 2005, the project has had an immediate and
staggering impact on energy consumption and utility costs. Space
comfort has improved, including the Data Rooms’ ability to maintain
temperature and humidity with in specified limits. The project
will conserve approximately 1 million KwH during the first (partial)
year in fully automatic operation. The Kw demand has almost
been cut in ½.
that only the HVAC system was optimissionedSM,
and that lighting and approximately 400 HP of hydraulic pumping
(serving the simulators) capability was left untouched, these
results are nothing short of spectacular. Due to the concerted
efforts of the optimissioningSM team, and especially Mr. Jeff
Seaton, approximately 2/3 of all HVAC equipment was turned off,
thereby generating these savings, extending the life expectancy
of the existing equipment, and reducing their associated maintenance
requirements and costs.
following charts were created from utility data provided by
Mr. Seaton, and generated by his Utility Manager software: