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Multi-Disciplinary Engineering Design Conference
Kate Gleason College of Engineering
Rochester Institute of Technology
Rochester, New York 14623
Project Number: 08452
INTEGRATION OF A RECIPROCATING COMPRESSOR
WITH A NON-TRADITIONAL RESEARCH
ENVIRONMENT
Garry Studley
Mechanical Engineer
Project Manager
Christopher Neitz
Computer Engineer
Interface Design
David Rigolo
Mechanical Engineer
Lead Engineer
James Jarvie
Industrial Systems Engineer
Integration
ABSTRACT
The objective of this project was to safely install a
reciprocating air compressor at a university, which required test
bed redesigns and interface development for future compressor
research projects. The compressor needed static and dynamic
structural support, as well as floor mounts for vibration
isolation. Compressor stage removals were designed to reduce
the required inputs and resulting outputs of electrical and
mechanical power, cooling fluids, and compressed air. An
intuitive graphical user interface was developed for use by
future engineering student teams. Potential future compressor
research projects include fault detection, variable valve timing
mechanisms, and auxiliary cooling units.
INTRODUCTION
In a demanding educational atmosphere, where universities are
attempting to differentiate their programs to add value for
prospective students, exciting projects need to be offered to
give today’s engineers significant real world experience. To
provide this, a technical university aligned itself with an
industry leader in the field of air compressors to create a variety
of unique projects that would highlight multiple facets of the
engineering curriculum. With this in mind, an air compressor
was chosen that would allow for multiple generations of
projects that could include interests in Mechanical, Electrical,
Industrial Systems, and Computer Engineering.
The selected air compressor was a vertical reciprocating,
multiple stage machine with positive displacement that was
used onboard navy aircraft carriers during the late 20th century.
The main function of the unit was turbine and diesel starting,
weapon augmentation, and to deliver liquid oxygen to storage
reservoirs that would be used for pilots while in flight. The
compressor was originally designed to compress the inlet
Kiernan French
Mechanical Engineer
Safety
condition of atmospheric pressure to a desired outlet pressure of
3000 psi (20.68 MPa) through six stages of compression
powered by a 75 HP (55.95 kW) electric motor. Designs for
the compressor required a compact unit that would be bolted to
the ship’s deck and make use of the abundant seawater for
cooling of the air in heat exchangers after compression.
Although this model of compressor is being removed from use
on aircraft carriers, it still provides a valuable test bed that
directly relates to products in the air compressor industry. The
unit could allow for implementation of various research
projects such as variable valve timing and fault detection that
are of particular interest to the industry. Value is also seen
from the educational standpoint as future engineers involved in
these projects explore areas of thermodynamics, vibrations,
data acquisition, control systems, and multiple other fields.
To effectively use the compressor in a university environment,
the machine needed to be transformed into an adequate test bed
that could be installed, revamped, and have an interface in
place to facilitate future research. Major challenges included
minimizing effects of the compressor on the educational
atmosphere, and accounting for future projects involving the
machine. Safety was the main concern due to the nature of the
compressor’s new environment that lacked the mechanical and
electrical resources and personnel with technical knowledge of
the unit for proper operation and maintenance. The compressor
input and outputs had to be modified in the interest of safety
while considering future project intentions in all redesigns of
parts or systems. The end goal of this project was to
successfully provide instructions, part designs, and a functional
interface that would create an appropriate test bed for future
teams.
NOMENCLATURE
© 2008 Rochester Institute of Technology
Proceedings of the KGCOE Multi-Disciplinary Engineering Design Conference
I – Electrical Current (A)
n – Number of Stages
 – Density (lb/ft3)
k – Ratio of Specific Heats

m - Mass Flow Rate (slugs/s)
PT – Mechanical Power (HP)

q - Heat Rate (Btu/s)T – Change in Temperature (°F)

V – Voltage (V)
V – Volumetric Flow Rate (ft3/s)
Z – Compressibility
rp – Overall Compressor Pressure Ratio
Cp – Specific Heat at Constant Pressure (Btu/lb*°F)
SAFETY
Throughout the project, safety has been a major consideration
at all levels. The main documentation for the project, the
Operation and Maintenance (O&M) Manual, was developed to
ensure that no one was injured while working on or operating
the compressor unit. Contained within the O&M Manual were
detailed instructions for safety considerations, as well as
sections for compressor startup, operation, shutdown, and
service. Additionally, a safety training session was developed
by the sponsoring company to provide all future teams with a
pre-project safety overview.
Safety was also a large consideration in the design and layout
of the test chamber. Structural, electrical, thermal,
environmental, acoustical, vibrational safety and fire hazards
were all considered. In order to mitigate electric shock hazards,
all electrical components were supplied by panels possessing
lock out-tag out points that would ensure the system was fully
de-energized when appropriate. Additionally, the electrical
system was designed to include two emergency stop buttons,
located on opposite ends of the test cell, which would
completely disconnect all electricity to the compressor. A
clearly marked fire extinguisher was located within the test cell,
with additional extinguishers located within the adjoining
machine shop. Warnings, cautions, and hazard signs were
placed throughout the test cell. These signs included; shock,
pinch, pressure, burn, and noise warnings. Ventilation within
the cell was checked to ensure that heat and exhaust gasses
would not make the room hazardous.
In order to ensure personal safety, requirements were developed
for all teams working on the compressor. The main requirement
was that the compressor be located in a test cell with restricted
access. This was achieved by ensuring that the test cell door
remained locked, and that anytime it was accessed, a log sheet
was signed indicating who was there and what was done with
the compressor. Requirements for personal protective
equipment were also developed. Safety glasses and hearing
protection were provided and are required whenever the
compressor is in operation, and the sound reducing doors must
be closed. When performing work on the unit, materials such as
work gloves, safety glasses, pants, and steel toed boots are
required. Additionally, every time a student team member
works on the compressor in the test cell, a faculty member or
lab supervisor must be informed.
INSTALLATION
The objective of the installation was to integrate the compressor
into its surroundings without adversely affecting the current
state of the atmosphere. The ability of the installation area to
withstand this type of compressor and resulting loads had to be
verified. Static and dynamic loads were a main concern as the
compressor would be located in a high traffic area of the
university. In addition to handling the loading of the machine,
the test cell had to provide the necessary inputs into the
machine and withstand the resulting outputs.
Proper
accessibility and maintenance capabilities were also required to
ensure safe operation of the compressor. Every aspect within
the installation had to consider the manner in which the
compressor would be used in the future to guarantee that the
proper test bed was available.
Structural:
The area that the compressor was scheduled to be installed in
was located between two precision machine shops, and it had
been previously used as an engine test cell. Although the room
was designed to be explosion proof, structural concerns existed
within the area. A main source of concern was the presence of
a basement located directly underneath the test cell. This
resulted in the floor thickness only being 4.75 inches (12.07
cm), which was supported by 12”x17” beams and structural
columns that created bays, area on the floor bounded by four
columns. Both the columns and beams are made out of
concrete with rebar reinforcement. The beams were oriented
across the base of the floor with a distance of 8.75 feet (2.64
meters) between the centers of adjacent beams. The overall
floor structure in that location of the building was designed for
80 lb/ft2 (3.83 kPa) of live load. Due to the floor condition, the
distribution of the 7500 lbs (33.36 kN) for the unit over the
20.49 ft2 (1.90 m2) footprint area of the machine was a concern.
Caution had to be used when developing a procedure for the
installation and placement of the compressor.
To gain an accurate estimate of the floor capabilities, a
professional structural engineer was hired to analyze the room
under the static and dynamic conditions that were imposed onto
it by the compressor. Based on the footprint area and the
weight of the compressor, a surface load of 372 lb/ft 2 (17.81
kPa) would be imposed onto the floor. After completion of the
analysis, the determination was made that the compressor could
only be placed onto the floor slab if the live load was limited to
50 lb/ft2 (2.39 kPa) in the bay location of the compressor. This
was based on the original structure’s live load capabilities being
exceeded with the placement into the room. With the presence
of the machine shop containing heavy manufacturing
equipment surrounding the test cell and the original capabilities
of the room, the live load could not be limited and the slab
would not be capable of supporting the load. Two steel support
beams were necessary underneath the footprint of the machine
to help the floor support the load from the compressor by
adjusting the effective bay size. With the installation of the
steel beams underneath the floor as opposed to a raised slab
under the footprint of the compressor, trip hazards were
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Proceedings of the KGCOE Multi-Disciplinary Engineering Design Conference
avoided. The weight of the compressor should be distributed
over 60 ft2 (5.57 m2) to alleviate the load on the floor during
transportation of the compressor to the test cell.
evenly distribute the weight of the machine for balance and
provide only an additional 6 inches (15.24 cm) in height.
Vibration Mounts:
Placement of the Compressor:
The physical characteristics of the compressor posed a difficult
scenario for the placement of the compressor into the test cell
location. The weight and dimensions of the unit were at the
brink of the test cell capabilities. This was due to the
transportation process involving the compressor being fully
assembled. The university did not have the resources necessary
to completely assemble the compressor. Careful examination
was necessary to ensure safety was upheld throughout the entire
process.
Clearance capabilities of the area were also
considered due to the compressor’s size relative to the test cell.
As was the case in examining the structural capabilities of the
test cell, the floor structure and underneath basement were
considered while the compressor would be transported into
place. Except for the initial drop off location of the compressor
at the machine shop loading bay, the transportation route would
travel across a floor structure, with basement underneath
similar to the test cell setup, for approximately 135 feet (41.15
meters). Under these circumstances, the additional weight from
the transportation equipment needs to be kept to a minimum. A
forklift could not be used due to this reason. A suitable pallet
jack that could safely transport the extensive weight of the
compressor was not available. The simultaneous use of two
pallet jacks was not a viable option due to safety concerns.
Experience and available resources at the university with
moving heavy equipment are limited, which led to the
requirement to hire a professional rigging company to perform
the transportation of the compressor.
The use of a professional rigging company provided the
experience and tools to also handle clearance and access issues.
Upon arrival at the loading bay, the compressor would need to
be moved through two doorways until it arrived in the test cell.
The clearance that the compressor will have moving through
the doorways needed to be determined. The compressor has a
width of 68 inches (1.73 meters), depth of 50 inches (1.27
meters), and a height of 79.88 inches (2.03 meters). The width
of the doorways did not pose a problem as the minimum
dimension, located at the test cell door, was 60 inches (1.52
meter) wide. The compressor could easily be oriented so that
the depth would be in that direction. The height clearance was
the primary concern, because each doorway to be passed
through only has a height of 83.5 inches (2.12 meters). With
the need to lift the compressor off the floor while transporting it
into location, the top 6 inches (15.24 cm) of the machine have
to be removed to reduce the overall height of the unit to 73.88
inches (1.88 meters). This portion of the compressor could be
placed back on the machine after it was in the test cell. While
extra clearance was added with this action, the rigging
company still needed to use the proper equipment to reduce the
lift off the floor to about 9 inches (22.86 cm). Based on the
weight and clearance issues, the rigging company d ecided that
utilization of four roller carts underneath the compressor will
In order to reduce the impact of the compressor’s operation on
the surrounding machine shop and classroom environment, it
will be mounted on a dampening system. This system will
reduce the amount of vibrations transmitted to the floor.
The dampening system acts to allow the machine a certain level
of acceptable motion. The majority of the vibrations produced
by the compressor are in the vertical direction, due to the
configuration of the pistons. With vibration isolators installed
underneath the compressor, the vertical oscillating load will be
dissipated by working to compress and depress the mounting
system. Essentially the force will be transferred into moving the
machine instead of the floor.
Caution must be taken when choosing the spring rate of the
mounting system, because the machine runs at a fixed level of
revolutions per minute (RPM). If the frequency of the
vibrations is within about 50 to 150% of the natural frequency
of the system as defined in Eq. 1, the vibrations will actually be
amplified by the mounting system. The desired ratio between
vibration frequency and natural frequency should be at least
1.5. To reach this ratio, the resonant frequency must be briefly
passed through as the compressor spins up and stabilizes; it is
not expected that this short time period of matching the
resonant frequency will cause significant effects.
Natural Frequency(rad / s) 
mount stiffness ( N / m)
Machine Weight (kg)
(1)
Originally four types of mounts were considered. These were:
mats of dampening material, sandwich type mounts, air filled
type mounts, and spring type mounts. The dampening mats
were ruled out early, despite their ease of installation and
reduction of points loads, because of their weak dampening and
high price. Sandwich type mounts were eliminated because,
when coupled with the compressor, they would produce a
spring mass system that would too closely resonate with the
compressor’s current designed RPM level. Air type mounts
were robust, but discovered to be more expensive than
originally anticipated. Considering price, deflection and spring
rates, and minimization of resonating frequencies, spring type
mounts provided the most appropriate solution.
Twelve mounts utilizing existing mounting holes around the
compressor perimeter would be used. The twelve out of twentytwo available holes were selected to achieve equal weight
distribution on the floor.
The original in-situ design of the compressor disregarded the
machine vibration impact on the surrounding environment,
because the compressor was designed to be rigidly mounted to
the deck of a ship using twenty-two 1.125 inch (2.86 cm) bolts.
To utilize these existing mounting holes on the base of the
compressor, some form of adaptation to the mount’s 0.5 inch
(1.27 cm) bolts needed to be made. For this purpose 0.25 inch
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Proceedings of the KGCOE Multi-Disciplinary Engineering Design Conference
(0.64 cm) thick 3 inch (7.62 cm) squares of steel plate will have
holes drilled in them to act as reducing washers.
To prevent the compressor from walking across the floor, the
mount would be anchored to the concrete floor. For this
purpose there is a wide selection of available anchors.
Vibration-resistant wedge stud anchors have been chosen due
to their healthy balance between high strength, minimal hole
depth requirement, and low price. A contracting company will
need to be hired to locate the rebar in the concrete floor prior to
the drilling of any holes during the mount installation. This
will prevent damage to the rebar in the concrete floor during
drilling procedures that would physically weaken the floor
support structure.
motor of the compressor would not be removed while in the test
cell. Removal of the motor would only be necessary upon
failure and require equipment not available at the university.
After subtracting the width of the compressor, the remaining
space available in the shorter direction of the room was 5.813
feet (1.77 meters). This space was split to give the operation
side of the compressor the maximum amount of space possible.
The position of the compressor in the long dimension of the
room was determined based on the need for storage and future
project uses of the machine. Shelving, a computer with a desk,
and workbench were all considered with the test cell layout.
The selected configuration of the room can be seen in Fig. 2.
Other hardware includes washers and nuts for the anchor studs,
and a level adjusting nut on the mount as shown in Fig. 1.
Machine Bolt
1/2-13 x7.5"
Adaptor plate
Compressor
frame
½” Anchoring
hardware
Figure 2: Selected Test Cell Layout
½” Flat
washers
Nut supplied
with mount
Spring mount
Figure 1: Vibration Mount Bolt Down Diagram
Because 0.5 inch (1.27 cm) bolts are being used in the 1.188
inch (3.02 cm) frame holes, a tolerance stack up showed that
there is room for up to 0.40 inches (1.02 cm) of error in the
placement of the mounts during installation into pre-drilled
holes in the floor.
Test Cell Layout:
Once the professional rigging company maneuvers the
compressor into the test cell, precise placement of the unit will
be needed to provide ample access area around the machine.
The test cell needed to be arranged to provide the proper access
area for maintenance capabilities and safe operation. The room
dimensions are 9.98 feet (3.04 meters) wide and 24.375 feet
(7.43 meters) long. Enough room for spare equipment and
storage area is also needed for proper maintenance and future
uses of the compressor.
To provide ample room for maintenance on both sides of the
compressor, the unit should be oriented near the center of the
room. Based on the dimensions of the compressor and room,
the depth of the machine was oriented to be with the width of
the room. This maximized the distance on either side of the
compressor. The minimum space needed for access on either
side of the machine was 23 inches (58.42 cm), assuming the
COMPRESSOR INPUTS AND OUTPUTS
With the insertion of the compressor into the university
environment, the inputs and outputs of the machine needed to
be altered to and verified that an appropriate level was
achieved. With the original design of the compressor, the
university would not be able to support the unit. Inputs such as
electrical supply, mechanical power, and cooling specifications
were not readily available due to lack of resources. Vibrations,
acoustics, pressure, heat generated, and other various outputs
imposed a threat to the general safety of the surrounding
environment.
With this in mind, the compressor was
redesigned to have reduced inputs and outputs by transforming
the unit from 6 stages with an exit pressure of 3000 psi (20.68
MPa) into 2 stages that would have a reduced overall discharge
pressure of 100 psi (0.69 MPa). With the redesign, the effects
on the inputs and outputs were determined to distinguish
whether further adjustments to the compressor or room were
necessary.
Mechanical Power:
By redesigning the compressor to operator with 2 stages instead
of the original 6 stages, the mechanical power needed by the
unit would be less than the 75 HP (55.95 kW) that was
originally supplied to the electric motor. Under the assumption
of an adiabatic cycle, the mechanical horse power required
could be determined based on the pressures and properties for
the compressed air as shown in Eq. 2[1]. The mass flow rate,

m , through the compressor was considered to be constant.
Page Number 4 of 8

k 1
n * p1 *V 1
Z  Z2
k
PT 
*
*r k * 1
229
k 1
2 * Z1
(2)
Proceedings of the KGCOE Multi-Disciplinary Engineering Design Conference
Equation 2 is a derivation of adiabatic horsepower using
English units. A conversion factor of 229 allows use of a
volumetric flow rate in gallons per minute. The input and
output conditions are represented by the subscripts 1 and 2,
respectively.
The subscript T denotes theoretical. The
theoretical horse power for the original design was computed
with the result from Eq. 2 being divided by the actual measured
horse power to estimate the compressor efficiency.
Calculations were then preformed, similar to the calculations
for the original design, to attain the theoretical horsepower for
the redesigned 2 stage compressor configuration. This result
was then divided by the compressor efficiency to determine the
estimated horse power the new design would require.
Numerical values used in the equation can be seen in Table 1
and Table 2.
Table 1: Properties of air used in Eq. 2 for each design
Table 2: Numerical values used in Eq. 2 for each design
The computed impact of the compressor redesign on actual
mechanical power was a reduction from 68 HP (50.73 kW) to
25 HP (18.65 kW). Specifications on the electric motor
indicated that the motor would be able to handle a reduced load
while maintaining a desirable motor efficiency. With reduced
power requirements, a safer environment could be maintained
and other inputs could also be reduced.
Electrical Supply:
One of the major advantages of reducing the mechanical
horsepower would be the effect on the electrical requirements.
The electric motor was determined to be capable of being
efficient at a reduced load which in turn would require less
electrical current draw by the motor.
The original design of the compressor called for 440 V
delivered at 90.3 A. While the university had the voltage
supply available at the test cell location, the current originally
specified would require a customized route to the location from
the main electric control panel through thick slabs of concrete.
This type of current would also require high capacity wires that
would significantly add to the cost. Due to the impact of the
redesign on the necessary power drawn by the motor while
continuing to use a 440 V supply, the current could be reduced
and still provide an ample supply of power. Equation 3[2] was
used to compute the new current requirements based on the
reduced horse power, compressor efficiency, motor efficiency,
and power factor associated with the motor shown in Table 3.
V * I * Power Factor 
HP * Compressor Efficiency * 746
motor efficiency
(3)
Table 3: Numerical values used for Eq. 2 for each design
The motor specifications table provided the necessary
information on the motor characteristics and the results from
computing the reduced horse power were used to gain the new
current requirement of 37 A. This result was verified by
examining the load percentage the motor would be at with the
new conditions and interpolating the resulting current value
from the provided information on the motor drawing. Original
startup of the compressor was in an unloaded state and the
necessary current used for this process was not a concern. The
new current requirement would allow the unit to be supplied
with the existing capabilities of the test cell.
Heat Generation and Removal:
Much like the electrical requirements, the requirements
concerning the heat released by the compressor were
anticipated to be reduced by the redesign with the reduction of
stages. Original specifications considered the compressor’s
application on a ship and utilized the vast supply of sea water
for an open loop, inner-stage cooling system.
With the removal of 4 stages on the high pressure side of the
compressor, the heat generated from compression was expected
to be reduced. This would result in lower requirements for the
open loop cooling system, which was originally designed for 50
gal/min (189.27 liters/min) of sea water to be used in the heat
exchangers between stages. In the redesign, the open loop
system had to be converted to fresh water, due to its availability
and ability to be a suitable alternative. Using an energy balance
assuming constant specific heat at constant pressure shown in
Eq. 3 and known temperature measurements at the inlet and
outlet of stages, the total heat removal of each stage could be
computed and summed together. Equation 4[3] could also be
used to calculate the actual heat removal being performed by
the open loop system with temperature measurements taken at
the beginning and end of the system.


q  m C p T
(4)
With the assumption that adiabatic inner-stage cooling was
occurring by way of the heat exchangers and a constant mass
flow rate of cooling water through the machine, theoretical
values for heat removal could be computed for the original and
redesigned compressor configurations, with 6 and 2 stages
respectively. The theoretical heat removal could then be
compared to the actual heat removal of the open loop system
similar to the method used in calculating compressor efficiency.
The inefficiency in the cooling system found from this ratio
could be used to predict the actual heat removal for the new
system.
With the actual heat removal of the redesigned
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Proceedings of the KGCOE Multi-Disciplinary Engineering Design Conference
compressor calculated, the required volumetric flow rate of the
system could be determined using Eq. 4 and the same
temperature difference that was measured for the original
design. The new required volumetric flow rate of the open loop
system to maintain the same temperature difference was
computed to be 18 gal/min (68.14 liters/min). Based on the
capabilities of the test cell, the tolerated temperature differential
in the system was raised slightly, from 4°F (-15.56°C) to 10°F
(-12.22°C), for a required flow rate of 7 gal/min (26.50
liters/min).
Another concern was the release of heat from the operation of
the compressor into the test cell room. With the doors closed
and no air leaving the room, the temperature would rise
significantly. To eliminate this effect, ample exhaust was
necessary to remove the air from the room. The current setup
in the test cell had a gravity inlet vent and two exhaust fans.
The total circulation that resulted from the exhaust capabilities
was 900 ft3/min (25.49 m3/min) and was more than sufficient
for the intended use.
and 3rd stage pistons with long shafts. Screwed into the top of
these pistons are shafts that lead to another set of pistons for the
4th, 5th, and 6th stages (refer to figure 3 for a visual
representation.) This allows the compressor’s reciprocation be
modeled the same as a three cylinder engine. This combined
with the varying weights of the pistons leads to a considerable
amount of unbalanced forces between the farthest and closest
pistons. The manufacture has provided us data on the
unbalanced inertia forces the machine transfers into the
foundation.
A silencer will also be used to further alleviate the impact from
the acoustics on the environment. The silencer will be placed
on the inlet of the machine and oriented in the vertical
direction. Due to an existing overhead light and the added
height from the silencer, interference issues arose. Options to
eliminate interference included moving the light to the side of
the compressor or installing an elbow at the inlet to redirect the
silencer past the light. The latter was chosen due to its overall
ease.
1st
piston
2nd
piston
Acoustics:
The reciprocating attribute of the compressor provided concern
about the effect it might have on the university environment.
Much of this concern was directed at the additional noise
related to the compressor’s operation. Verification that the
educational atmosphere would be preserved was necessary.
Aspects of the room layout, room capabilities, and machine
characteristics were considered for this purpose.
Features of the test cell, such as being sealed with an acoustic
reducing steel door, were anticipated to help the sound issue
brought about by the compressor. The test cell capabilities to
mitigate any added noise levels were tested by a high volume
impulse generating device. Measurements were taken with a
decibel meter inside and outside the test cell with the door
closed to determine the overall differential. The results from
this test describe the capabilities of the room to dissipate sound
with an average differential of 64.8 dB. With the measured
acoustic levels from the compressor being no more than 90 dB,
the acoustic disturbance from the machine was viewed to be
minimal.
6th
piston
5th
piston
4th
piston
3rd
piston
Crossheads
Figure 3: Piston configuration
The 4th, 5th and 6th stage pistons will be completely removed.
Because of this, their weight will have to be compensated for.
The 3rd piston will remain in the compressor because there is no
benefit of removing it. To counter the removed weight, a donut
shaped weight will be bolted to existing holes on each of the
crossheads as shown in Fig. 4.
Shaft to 3rd and 4th pistons
Crosshead
REVAMP
The pistons act in a vertical arrangement and are of varying
weights. The stages are configured such that there are only
three rods connected to the crank that then connect to three
crossheads. These crossheads are then connected to the 1st, 2nd,
Page Number 6 of 8
Support surface
4th stage counterweight
Crank
Figure 4: Sample Counterweight Assembly
Proceedings of the KGCOE Multi-Disciplinary Engineering Design Conference
The largest of the three counter weights will be approximately
14 lbs (62.28 N). Figure 4 shows this weight disassembled
along with the rod-crosshead assembly. The front and back
edges of the crosshead act as a bearing surface. This surface
interacts with a cylindrical lining built into the block to support
the linear reciprocating motion of the crosshead. This particular
counter weight has wings sticking out of the side of it because
of clearance issues. The crosshead drops below a side
supporting surface (surface of crosshead that interacts with
support surface is labeled in Fig. 4) in the block prohibiting
material from overhanging over the top of the front and back of
the crosshead. Because of limited space above the crosshead,
the next feasible place to add mass would be to the sides. The
other two counterweights would not need this extra material,
because they are only 3 to 4 lbs (13.34 to 17.79 N), and could
fit in the foot print on the top of the crosshead. They are lighter
because the 5th and 6th stage pistons are very small. The
counterweight’s two pieces are designed to interlock to prevent
bending loads on the mounting bolts that would lead to failure.
With the removal of the 4th, 5th and 6th pistons, their cylinder
heads will also be removed. The 4th and 5th stage pistons are
connected to the 3rd and 2nd pistons respectively. The 2nd and 3rd
stage piston only act to compress gas on the downward stroke.
Figure 3 shows each pistons compression volume as a
crosshatched area for reference. Because of this and the fact
that the 4th and 5th stages are attached to the top of these
pistons, plates can be used to block off where the 4 th and 5th
cylinder heads and pistons where originally located. The 1st
stage piston acts on both the upward and downward stoke.
Therefore 1st stage has an upper cylinder head unlike the 2nd
and 3rd stages. Because of this, care must be taken to account
for the volume of the removed 6th stage components. For this
purpose a plate with a length of cylindrical bar will be used as
shown in Fig. 5. The bar will extend down slightly past where
the seal for the 6th stage piston rod is. This will effectively seal
the 1st stage while simulating the majority of the compression
volume of the 1st stage taken up by removed 6th stage
components.
6th stage block off plate
Fake 6th stage
Figure 5: 1st Stage Seal and 6th Stage Cover Plate
Other components being removed from the compressor
included the 3rd, 4th, 5th, and 6th stage air/gas coolers, water
separators, air piping, relief values, and coolant piping. None of
these items will be utilized since only the 1st and 2nd stages are
to remain in operation. The air piping is run in series between
the stages. Therefore the air piping after the 2nd stage relief
valve will be removed and a new pressure tap will be installed.
The open loop coolant piping runs in parallel through the six
air/gas to coolant heat exchangers. The 3rd through 6th stage
coolant pipes will be replaced with plugs installed in the inlet
and outlet coolant manifolds to take their place. All these
components will be stored on shelving for possible use in future
projects.
With only the first two stages in operation, the compressor will
compress air to 100 ±20 psi (0.69 ± 0.14 MPa). This is a much
safer pressure for an academic environment than 3000 psi
(20.68 MPa). Also the removal of stages reduces the required
motor power from 68 HP (50.73 kW) to only 25 HP (18.65
kW). This will lead to the electrical current requirement of the
motor dropping to 38 Amps from 90 Amps.
INTERFACE
Because the future goal of the compressor revamp is to allow
for users to monitor the state of the compressor during various
research projects, a data acquisition (DAQ) system needed to
be created. Using LabVIEWTM as a foundation, an interface
was created which can be used to build a full data acquisition
system once the compressor is installed and converted to a 2stage system.
The features of the interface were created with a primary focus
on user convenience to view, record, and trend data with ease.
The first function of the interface will gather data from sensors
to be installed on the compressor. Temperature and pressure are
common examples of data collected, but additional data such as
water flow and crank angle may be required in order to aide in
future research.. Currently, the interface is set to display 25
different sensor indicators, and additional sensors can easily be
added. The next important ability of the interface is to store the
data being displayed. In the event the user would like to take
certain ranges of points and trend them, a single file would
need to be created that could be read by data manipulation
programs such as Microsoft Excel. For real-time visualization
of data trends through the interface, graphs are included for
each sensor, which a user could utilize via a set of organized
tabs. A threshold line can be adjusted for each graph, so a user
can more clearly see how close the data is to an unsafe level.
Finally, the interface required a standardized warning system,
which could alert the user that a part of the compressor had
gone over a safe threshold level. This was accomplished via a
number of status diodes, which glow green while the data being
reported is under a safe level, but glow red when an unsafe
level is reached. Using this mechanism, a user could easily and
quickly see where on the compressor a problem was occurring.
Usability testing, using both objective and subjective data
collection methods, was implemented with a representative set
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Proceedings of the KGCOE Multi-Disciplinary Engineering Design Conference
of individuals for the intended user group. While being asked
to complete a set of specified tasks with the interface, objective
data was collected on the number of errors committed, the
number of times help was requested, and the time taken to
complete each of the tasks. Individuals repeated the test a
second time, and comparisons in the data were made to
determine the high degree of learn-ability present in the
program. The individuals then completed a brief survey, which
was used to capture the quality of subjective aspects such as
intuitiveness, aesthetics, and an opportunity to provide written
comments.
With these features, a simple but robust interface was created
which allowed a user to view and record data, monitor trends of
past data, and view whether or not the data was within a safe
limit. Fig. 6 shows a screenshot of the front panel of the
interface. Sensors are grouped together based on their function
and location. Graphs for each grouping are accessible with
localized buttons, as well as tabs positioned across the top of
the interface. Finally, the diode status lights provide a clear
indication on the overall health of the system.
revamp, and interface design still have valuable processes that
can be followed as a reference.
Installation:
The basic layout of the test cell room can stay the same in most
regards depending on size of the new compressor. All supplies
and equipment purchased for operation and maintenance can be
utilized with the new compressor. The structural analysis that
was provided for the room could be used as a benchmark for
the capabilities. The recommendations to strengthen the
structure with steel beams underneath the floor should be
followed with the addition of any new large equipment to the
intended location. A rigging company in the intermediate area
was specified to be scheduled to perform the installation.
Vibration mounts used for isolation have also been identified,
and will need to be ordered with a selected quantity as dictated
by the new compressor footprint.
Revamp:
While this portion of the project probably will not directly
relate to the new compressor, the basic ideas can be applied to
any reciprocating compressor. Theory and proper equations
have been documented to compute the resulting changes that
the revamp had on the machine, such as the reduced mechanical
power, reduced electrical requirements, vibration analysis, the
cooling capabilities needed with the inner-stage cooling, and
cooling of the room from heat generated during operation. The
theoretical calculation process can be followed to address the
effect the revamp would have on the new machine.
Interface:
This section can completely transfer over to the new
compressor. A working interface, a user guide, a method for
data collection, and a computer have all been provided.
Specifications on the DAQ hardware that should be used for
the future project’s use have been identified and recommended
to be ordered. The quantity of stages that the new compressor
will have should be addressed with an adjustment made in
ordering and display on the interface.
Figure 6: Front Panel of the Interface
CONCLUSIONS
Although the entire scope of the project was on course to be
completed, the compressor was never received by the university
due to issues with the International Traffic in Arms Regulation
(ITAR) control designation of the machine. This basically
meant that non-US citizens could not transport any information
regarding the compressor outside the workspace, due to the
compressor’s use on government ships. While certain non-US
citizens could still work on the compressor, a major liability
risk was present with the consequence of large fines if
regulations were violated. Extra security measures would have
also been required. The legal councils of both parties decided
against the donation for these reasons.
As a result, another compressor, also reciprocating, will be
selected and installed into the same area that the original
compressor was to be located. The future purposes of the new
compressor will be identical. With these similarities, a great
deal of this project can be used in the work intended for the
new compressor. The three main subdivisions of installation,
ACKNOWLEDGMENTS
Dresser-Rand Company, Rochester Institute of Technology
Jensen Engineering, Dr. Margaret Bailey, Mr. Mike Bunce, Mr.
Scott Delmotte, Mr. Dave Hathaway, Dr. Mark Kempski
Mr. Allan Kidd, Mr. Ray McKinney, Mr. Joe Tecza, Mr. Jason
Vigil
REFERENCES
[1] Ingersoll Rand 1981,Gas Properties and Compressor Data,
Ingersoll Rand Co. USA., pp. 2-20.
[2] Bodine Electric Co. 1993, Small Motor, Gearmotor and
Control Handbook, Bodine Electric Co., Chicago, IL.
[3] Moran M. and Shapiro, H 2004, Fundamentals of
Engineering Thermodynamics, John Wiley & Sons, Inc.,
Hoboken, NJ.
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