Energy Efficient Compressed Air Systems
Introduction
Compressed air systems generate, store and distribute energy in the form of compressed air for
use throughout a plant. In a compressed air system, a single set of compressors can supply
power to machines all over the plant, thus eliminating the need for numerous and dispersed
electric motors. This advantage must be balanced against the relative poor energy efficiency of
compressed air systems, which can be as low as 20% when leaks and part-load control losses are
taken into account.
On a national scale, air compressors rank only behind pumps in terms of industrial motor drive
electricity consumption. Thus, increasing the efficiency of compressed air systems can result in
significant energy savings.
Principles of Energy-Efficient Compressed Air Systems
Energy Balance Approach
To compress air, the power delivered to the fluid (air) dWf is the integral of the product of the
volume flow rate V and the pressure rise dP.
dWf = ∫ V dP 
The electrical power supplied to an air compressor is:
dWe = ∫ V dP / (motor compressor control)
where motor is the motor efficiency, compressor is the compressor efficiency and control is
the control efficiency.
Three types of compression are shown below. The right compression line represents isentropic
compression, in which air is compressed adiabatically with no internal reversibilities. The left
compression line represents isothermal compression, in which the air is cooled to keep the air
temperature constant during compression. Isentropic compression has no cooling and
isothermal compression has the maximum cooling possible. Actual compression processes lie
somewhere in between isentropic and isothermal compression, and are called polytropic
compression. The area to the left of the compression lines represents the fluid work dWf = ∫ V
dP. Thus, isothermal compression requires less compressor work because the cooling is
responsible for part of the decrease in volume.
Energy Efficient Compressed Air Systems
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Source: Cengal, Y. and Boles, M., Thermodynamics, 1998, WGB-McGraw-Hill.
Some air compressors utilize two stages of compression with intercooling between the stages to
further reduce compressor power. The power savings from two-stage compression with
intercooling are shown graphically below.
Source: Cengal, Y. and Boles, M., Thermodynamics, 1998, WGB-McGraw-Hill.
Assuming that air can be treated as an ideal gas, it can be shown that
Pvn = constant
during the compression process, where P = absolute pressure, v = specific volume, n = 1 for
isothermal compression, n= k = Cp/Cv = 1.400 for isentropic compression of air and 1.0 < n <
1.400 for polytropic compression.
Substituting (Pvn = constant) into the equation for fluid work (dWf = ∫ V dP) and solving the
differential equation yields the following results:
Wf = R T ln (P2/P1) for isothermal compression
Wf = n R T1 [(P2/P1) (n-1)/n - 1] / (n - 1) for polytropic compression
Energy Efficient Compressed Air Systems
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Rair = 0.06855 Btu/lbm-R
Example:
Calculate specific capacities (cfm/hp) for isothermal and isentropic compression of 70 F air to 100
psig.
Fluid work for to compress 70 F air to 100 psig isothermally
Input Data:
R (Btu/lbm-R)
0.06855
T®
530
P1 (psia)
15
P2 (psia)
115
p at std cond (lb/ft3)
0.075
Calculations:
Wf = R T ln (P2/P1) (Btu/lbm)
74.003
Wf (hp/cfm)
0.131
1 / Wf (cfm/hp)
7.642
Fluid work for to compress 70 F air to 100 psig isentropically
Input Data:
R (Btu/lbm-R)
0.06855
T®
530
P1 (psia)
15
P2 (psia)
115
p at std cond (lb/ft3)
0.075
n
1.4
Calculations:
Wf = n R T1 [(P2/P1) (n-1)/n - 1] / (n - 1) (Btu/lbm)
100.400
Wf (hp/cfm)
0.178
1 / Wf (cfm/hp)
5.633
Actual compressors generate between 4 and 5 scfm/hp at 100 psig. The difference between the
thermodynamic values of scfm/hp computed above and scfm/hp generated by actual
compressors is due to the turbulence and friction generated within the compressor. Thus, this
difference characterizes the efficiency of the compressor.
Example:
Calculate the efficiency of a compressor with an actual specific capacity of 4.2 cfm/hp if the
polytropic specific capacity is 6.0 cfm/hp.
dW = ∫ V dP / compressor
compressor = ∫ V dP / dW
compressor = 4.2 scfm / 6.0 scfm = 70%
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Motor efficiency is the efficiency of the motor at converting electrical power into shaft power.
The efficiency of a premium-efficiency 100-hp motor is about 92%. Motor efficiency can be
improved by specifying premium-efficiency motors.
Control efficiency is a measure of the losses incurred to vary compressed air output to match
compressed air demand. In air compressors, control efficiency varies widely depending upon the
type of part-load control employed.
Understood in this light, the energy balance equation serves as a useful guide for energy saving
opportunities. Thus, primary energy savings opportunities are:





Reducing volume flow rate
Reducing pressure rise
Increasing control efficiency
Increasing compressor efficiency
Increasing motor efficiency.
Opportunities for Improving The Energy-Efficiency of Compressed Air Systems
These principles can be organized using the inside-out approach, which sequentially reduces enduse energy, distribution energy, and primary conversion energy. Combining the energy balance
and inside-out approach, common energy-efficiency opportunities in compressed air systems
include:



End use
– Eliminate inappropriate uses of compressed air (reduce V)
– Install solenoid valves to shut off unnecessary air (reduce V)
– Install air saver nozzles (reduce V)
– Replace timed-solenoid with differential-pressure control (reduce V)
– Use blower instead of air compressor for low-pressure applications (reduce dP)
Distribution
– Fix leaks (reduce V)
– Replace timed-solenoid drains with demand-control drains (reduce V)
– Decrease pressure drop in distribution system (reduce dP)
Conversion
– Compress cooler outside air (increase compressor efficiency)
– Stage compressors with pressure settings or controller (increase control efficiency)
– Employ on/off, load/unload with auto shutoff, or variable-speed control for trim
compressor (increase control efficiency)
– Add compressed air storage to decrease unload power and increase auto-shutoff
(increase control efficiency)
– Replace desiccant with refrigerated dryer (reduce V)
– Use heat from compressors to heat building during winter
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Recurring Energy-Efficiency Concepts
Close inspection of these energy-efficiency opportunities illustrates three important and
recurring energy efficiency concepts.

The equation for air compressor energy use serves as a useful guide for comprehensively
identifying energy saving opportunities.

Like most systems, compressed air systems are designed for peak conditions, but spend
the vast majority of time operating at off-peak conditions. Thus, several energy efficiency
opportunities result from improving control to reduce unnecessary compressed air use
and power consumption during off peak conditions. Careful attention to “control
efficiency” is vital to achieving energy efficiency.

To achieve energy savings, many end-use and distribution system savings opportunities
must be coupled with modifications to the conversion equipment, which in this case is the
air compressor plant. Thus, the ‘whole-system’ inside-out approach is vital to maximizing
energy-efficiency potential.
Air Compressors
The three basic types of air compressors are reciprocating, rotary screw and centrifugal
compressors.
Reciprocating
Rotary Screw
Centrifugal
Reciprocating compressors use pistons to compress air in cylinders. Single-acting compressors
compress air on one-side of piston, and double acting compressors compress on both sides of
piston. Large reciprocating compressors may employ multiple stages with intercoolers and
double acting pistons to achieve high compression efficiencies. Single-stage compressors control
compressed air output by stopping the pistons when compressed air is not needed. Multi-stage
compressors control compressed air output by sequentially reducing the number of stages in use.
Energy Efficient Compressed Air Systems
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Rotary-screw compressors compress air by forcing air between rotating screws with decreasing
volume between the screws. Most rotary-screw compressors control compressed air output by
modulating the air intake valve, and or alternating between full open and fully closed operation.
Centrifugal compressors compress air by accelerating air from the tips of impellors rotating at
high speeds into a volute. Centrifugal compressors are typically 250-hp or larger, and frequently
employ multiple stages to achieve the desired compressed air output pressure. Centrifugal
compressors control compressed air output by modulating an inlet valve or variable inlet vanes
on the air intake, loading and unloading, or blowing off compressed air to atmosphere rather
than into the compressed air system.
Compressor Controls
Compressor controls typically match compressed air output to compressed air demand by
maintaining discharge air pressure within a specified range. There are five primary control
strategies for maintaining the pressure within the desired range.
On/Off Control
In on/off control, the compressor turns on and begins to add compressed air to the system when
the system pressure falls to the lower activation pressure. The compressor continues to run and
add compressed air to the system until the system pressure reaches the upper activation
pressure when the compressor shuts off. Typical lower and upper activation pressures would be
90 psig and 100 psig. On/off control may also employ a timer to reduce short-cycling.
Reciprocating compressors typically employ on/off control. On/off control is the most efficient
type of part-load control, since the compressor draws no power when it is not producing
compressed air.
Load/Unload Control
In load/unload control, the compressor “loads” and begins to add compressed air to the system
when the system pressure falls to the lower activation pressure. The compressor continues to
run and add compressed air to the system until the system pressure reaches the upper activation
pressure. It then “unloads” and does not add compressed air to the system until the system
pressure drops to the lower activation pressure. Typical lower and upper activation pressures
would be 90 psig and 100 psig. When unloaded, rotary screw compressors typically partially
close the air inlet valve and bleed the remaining compressed air in the sump to atmosphere.
Power draw when fully unloaded varies from about 60% of full load power to about 30% of fullload power, depending on compressor design and on the length of time the compressor runs
unloaded. To fully unload, the load/unload cycle time must be long enough to allow the
compressed air in the sump to bleed to atmosphere when the compressor unloads. Thus,
load/unload control works best when coupled with adequate compressed air storage, which
lengthens load/unload cycles while modulating pressure variation to end uses.
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Most compressors with load/unload control also have an “automatic shutoff” option, in which
the compressor shuts itself off if it runs unloaded for about 5 to 10 minutes. The compressor will
remain off for a specified period of time before restarting to avoid short-cycling. Running the
compressor in “automatic shutoff” mode can result in significant energy savings during periods of
low compressed air demand. In addition, adequate compressed air storage increases
load/unload cycle time, and the likelihood that the compressor shuts off after running unloaded
for a few minutes.
Modulation Control
In modulation control, the position of the inlet air valve is modulated from full open to full closed
in response to compressor output pressure. Modulation control typically employs PID control
with a narrow control range about + 2 psig. Inlet modulation is a relatively inefficient method of
controlling compressed air output.
Variable-Speed Control
Rotary-screw air compressors can be equipped with variable frequency drives to vary the speed
of the screws and the corresponding compressed air output. As in other fluid flow applications,
the variation of speed to vary output is extremely energy efficient.
Blow-off Control
In centrifugal compressors, the quantity of air flow through the compressor can only be
controlled by modulating the inlet air valve over a relatively small range. When flow is reduced
below this range, the flow becomes unstable in a “surge” condition. To avoid surge, centrifugal
compressors may discharge compressed air to the atmosphere to control compressed air output
to the system. Blow-off control is the least efficient method of controlling compressed air output,
since input power remains constant as the supply compressed air to the system decreases.
Power / Output Relationships by Control Type
The following figure shows typical relationships between fraction input power to the compressor
(FP) and fraction compressed air output (FC) for various types of control. At full output capacity
(FC = 1.0), compressors draw full power (FP = 1.0). The power draw at less than full output
capacity is a function of the type of part-load control. The figure shows that at part load, most
energy efficient control mode is on/off, followed by variable speed, load/unload, modulation and
blow-off control.
Energy Efficient Compressed Air Systems
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1.00
Fraction Power (FP)
0.75
Blow Off
Modulation
0.50
Load/Unload
Variable Speed
On/Off
0.25
0.00
0.00
0.25
0.50
0.75
1.00
Fraction Capacity (FC)
Assuming linearity, fraction power, FP, can be calculated from fraction capacity, FC, and fraction
power at no load, FP0, according to the following relationship:
FP = FP0 + (1 – FP0) FC
Some compressors use a combination of basic control modes described above. For example, the
figure below shows the relationship between fraction of full-load power and fraction of full-load
output capacity for a compressor using a combination of modulation and load/unload control.
The top line shows full modulation control, in which the compressor continues to draw 70% of
full load power even when producing no compressed air. The bottom line shows a combination
modulation and load/unload control, in which compressed air output is modulated by the inlet
valve down to 40% of total capacity. Below 40% of full output capacity, the compressor loads and
unloads to vary compressed air output. In this example, the compressor draws 25% of full-load
power when fully unloaded.
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Centrifugal compressors typically employ three primary methods to control compressed air
output to meet demand: inlet modulation by a flow control valve or variable inlet vanes,
load/unload and blow off. Inlet modulation by a flow control valve or variable inlet vanes varies
the quantity or rotation of inlet air to the compressor, which reduces compressed air output and
input power. In most cases, however, the control range using inlet modulation is limited to
between about 70% and 100% of compressed air output. If flow is reduced below about 70% of
full output capacity, an unstable flow condition called “surge’ may result. To control flow below
about 70% of full output capacity, some compressors can load and unload to match compressed
air demand. When fully unloaded, the compressor generates no compressed air and can draw as
little as 15% of full load power. Centrifugal compressors without load/unload capability continue
to generate compressed air, but blow off the excess compressed air to the atmosphere. Because
compressor power remains constant while compressed air output falls, blow-off control is the
least efficient method of controlling compressed air output.
Many centrifugal compressors employ some combination of these basic control options. For
example, the figure below shows the fraction power to fraction capacity curves for a centrifugal
compressor with “Constant Pressure” and “Auto-dual” control modes. In “Constant Pressure”
mode, variable inlet vanes modulate inlet air to the compressors down to about 70% of full load
capacity, and compressor power draw follows linearly. If compressed air demand falls below
70%, blow off valves discharge compressed air to the atmosphere and power draw remains
constant. Alternately, the compressor could be set to run in “Auto-dual” mode. In Auto-dual
mode, the variable inlet vanes modulate inlet air to the compressors down to about 70% of full
load capacity, just as in Constant Pressure mode. However, in Auto-dual mode, the compressor
will unload when compressed air demand falls below 70% of full-load capacity and compressor
power draw will be reduced to about 15% of full load power. The plot below shows fraction of
full load power draw (kW) on the vertical axis and fraction of full load capacity (cfm) on the
horizontal axis for Constant Pressure and Auto-dual modes.
Energy Efficient Compressed Air Systems
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1.0
0.9
0.8
Fraction Power
0.7
0.6
Constant Pressure
0.5
Autodual
0.4
0.3
0.2
0.1
0.0
0.0
0.1
0.2
0.3
0.4
0.5
0.6
0.7
0.8
0.9
1.0
Fraction Capacity
Lubrication
In reciprocating and flooded-screw compressors, lubricating oil comes in direct contact with the
compressed air. Most the oil is removed from the compressed air stream in the sump and by a
separator. However, trace amounts of oil are carried forward in the compressed air. In oil-free
screw and centrifugal compressors, no oil comes in contact with the compressed air.
Cooling
The temperature of air increases during compression and from irreversibilities within an air
compressor. Removing heat during compression reduces the work required to raise the pressure
of the air.
Heat can be removed from the air compressor to the surrounding air or to water. Air-cooled
compressors pass the hot lubricating oil from the compressor and compressed air through finnedtube heat exchangers and force ambient air across the heat exchangers using a cooling air fan.
Cooling fan horsepower is typically about 5% of the power of the compressor motor. Watercooled compressors use water-to-air heat exchangers to remove heat from the lubricating oil and
compressed air. In many applications, this heat is eventually rejected to the atmosphere by a
cooling tower. Increasing cooling by decreasing the temperature of the cooling air or water
improves compressor efficiency and output capacity.
In many cases, the relatively low temperature of the cooling air or water leaving an air
compressor limits its usefulness for other process-heating applications. However, the
temperature is typically high enough to provide useful space heating during winter. Thus, an
excellent application for reclaiming heat from air compressors is to direct hot air into the plant
during winter. About 70% of the electrical power input to an air compressor is typically removed
by the cooling system.
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Compressed Air Storage Tanks
Stored compressed air dampens variation in supply and demand and stabilizes compressed air
pressure within the system. Primary compressed air storage refers to tanks located near the air
compressor that dampen pressure variations for the entire compressed air system. Secondary
compressed air storage refers to tanks located near individual compressed air loads with variable
demand. The typical price for compressed air storage tanks is about $4 per gallon.
Primary compressed air storage tanks are generally sized to hold at least 10 seconds of
compressor capacity. For example, the recommended minimum volume of the primary storage
for a 25-hp air compressor generating 4 cfm/hp would be about:
25 hp compressor x 4 cfm/hp = 100 cfm x 10/60 min = 17 ft3
Primary storage tanks can be located upstream or downstream of the dryer. Locating the
primary storage tank downstream of the dryer reduces the variation in compressed air flow
through the dryer caused by large variations in end-use demand. In many cases, this eliminates
problems of excess water in the compressed air lines due to excess flow through the dryer during
periods of high compressed air demand.
Dryers
The two most common types of dryers for removing moisture from compressed air lines are
refrigerated dryers and desiccant dryers. Dryers are typically sized to handle the peak air
compressor air flow.
Refrigerated Dryer
Energy Efficient Compressed Air Systems
Desiccant Dryer
11
Refrigerated Dryers
Refrigeration dryers cool the air to a dew-point temperature of about 35 F and remove the
resulting condensate. According to manufacturer data, the typical power required for a
refrigerated dryer is about 6 W/scfm. For example, assuming a 200-hp compressor generates 4.2
scfm of compressed air per horsepower, the output capacity of the compressor at full load is
about 840 scfm. The power draw of a refrigerated dryer sized for this application would be
about:
840 scfm x 0.006 kW/scfm = 5.0 kW
Desiccant Dryers
Desiccant dryers adsorb water into desiccants and reduce the dew point temperature of the
compressed air to about -40 F or lower, which is much dryer than the compressed air leaving
refrigerant dryers. Desiccant dryers typically have two tanks. Compressed air flows upward
through the left desiccant tank where moisture is adsorbed by the desiccant (in an exothermic
reaction which warms the desiccant and air). After the desiccant becomes saturated with water,
the flow of compressed air is directed upward through the right desiccant tank. The left
desiccant tank is then purged of water.
Desiccant dryers employ three methods to purged water from the desiccant: compressed air
purge, heated compressed air purge and heated blower purge.
In compressed air purge, about 15% of the dry compressed air leaving the dryer is expanded to
about 45 psig and directed downward through the wet tank to purge moisture from the tank.
Example
Calculate the power requirement for compressed air purge for a 200 hp compressor
generating 840 scfm of compressed air if the compressor generates 4 scfm/hp.
Assuming the efficiency of the motor is 90%, the power required for purging the
desiccant in an unheated purge-type dryer would be about:
840 scfm x 15% / 4.2 scfm/hp / 90% x 0.75 kW/hp = 25 kW
25,000 W / 840 scfm = 30 W/scfm
This example shows that the drying power requirement or compressed air purge is about 30 W
/scfm. This is about five times as much electricity as a refrigerated dryer would use; thus,
desiccant dryers should be used only in applications that require very dry air.
In heated compressed air purge, about 7% of the dry air leaving the left desiccant tank is
expanded, then heated by electrical resistance heaters to about 375 F, and then directed
downward through the right tank to purge moisture from the tank. According to product
literature, heating requires about 7 W /scfm. Thus, the total power required for heated
compressed air purge drying is about 22 W /scfm.
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In heated blower purge, a blower supplies ambient air to an electric resistance heater that heats
the air to about 375 F. The heated air is directed downward through the right tank to purge
moisture from the tank. According to product literature, blowers require about 3 W/scfm and
heaters require about 13 W /scfm. Thus, the total power required for heated blower purge
drying is about 16 W /scfm.
The purge cycles are initiated by a timer or by a humidity sensor that determines the whether the
on-line tower has additional adsorptive capacity. Timers are generally set to handle peak
conditions when the air is the most humid. In many parts of the country, the summer air
contains 4 more times more humidity than winter air; thus, timed cycles typically use for more
energy for purging than is necessary during much of the year.
Condensate Drains
As compressed air cools, water vapor can condense out of the air and should be removed from
the compressed air system through drains. Condensate drains should be located:




After the after cooler
Underneath the receiver tank
At low points in the system
After filters, regulators and other devices that result in a large pressure drop.
Condensate is typically removed from compressed air systems by timed-solenoid or demandactivated drains. Timed-solenoid drains are controlled by a solenoid valve on a timer that opens
at a prescribed interval to discharge condensate. Although simple, reliable and inexpensive,
timed solenoid drains typically discharge excess compressed air because the timer is typically set
for peak condensate conditions during hot and humid summer months. Demand-activated drains
function like steam traps and remove condensate only when needed.
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If the condensate removal system does not function properly, cocks may be left partially open so
water doesn’t accumulate in the compressed air lines. This is a major waste of energy. The dryer
and condensate drains should be inspected, evaluated, and repaired or replaced.
Filters, Regulators and Lubricators
Filters, regulators and lubricators are frequently located in a grouped F-R-L sequence. The filter
removes particulates entrained in the compressed air and may have a trap or drain at the
bottom. Filter cartridges should be replaced when the pressure drop across the filter exceeds
about 7 psig. According to a compressed air service company, a filter cartridge for a filter on a 1inch line passing about 170 scfm costs about $30 and takes about 10 minutes to replace. Most
filters have float drains that open to discharge condensate when the condensate level reaches a
set point in the filter. The seats on these drains can become damaged or dirty allowing
compressed air to continually leak through the drain. According to a compressed air service
company, replacement float drain assemblies cost about $18 and take about 10 minutes to
replace. According to a compressed air service company, a good preventative maintenance
program would replace filter cartridges about every 1 year or 8,000 hours of service.
Regulators reduce downstream air pressure. Regulators have pressure gauges and valves to
adjust the downstream pressure.
Lubricators add lubricating oil. Lubricators look like filters, but have a clear bubble or screw
assembly on top for adding oil.
F-R-L groups are common locations for leaks and should be inspected regularly. In addition, if the
all machines in a plant or area use regulators to reduce air pressure, it would save energy and
compressor wear and tear to reduce the operating pressure of the compressor instead of
reducing line pressure with regulators.
Distribution System
The air distribution system includes headers, branch lines, hoses, valves and fittings. The
distribution system should be designed so that the total pressure drop from the compressor to
the farthest air-using machine is no more than 10 psi. Large headers serve both to minimize
pressure drop and increase storage. The most efficient layout utilizes a loop design for the
header pipe and a single compressor entry location. Typical rules for sizing compressed air
distribution lines are:
Main line: size from average cfm to get P < 3 psi
Branch line: size from cfm peak to get P < 3psi
Feed lines: size from peak cfm to get P < 1 psi
Hose: can generate P = 4 to 5 psi (proper selection of hoses is important!)
Energy Efficient Compressed Air Systems
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In addition, it is important to connect multiple compressors into the system using a piping design
which does not cause excess turbulence. For example, the diagram below shows 10 psig of
unnecessary pressure drop caused by a straight T-connection between compressors. The
pressure drop could be eliminated using curved or 45-degree connections.
Pneumatic Tools
Most pneumatic tools are designed to operate at 90 psig. Operating at a higher pressure
shortens tool life. Operating at a lower pressure may compromise the ability of the tool to
perform its task.
Low-Pressure Blowers
Low pressure blowers provide compressed air at pressures up to 20 psig using much less
electrical power than traditional air compressors, which generate 4-5 scfm/hp when compressing
air to 100 psig. For example, a positive displacement blower requires 43 bhp to provide 310 scfm
at 20 psig, and 17.7 bhp to provide 423 scfm at 5 psig. The The specific capacities (scfm/hp) at
these operating conditions are:
20 psig: 310 scfm / 43 hp = 7.2 scfm /hp
5 psig: 423 scfm / 17.7 hp = 24 scfm /hp
Compressor Sizing
The determine the required output capacity of a compressor, find the sum of the scfm
requirements of all the individual equipment, add 10% for leakage, and then size the compressor
to meet this load. As an example, consider the following case:
Energy Efficient Compressed Air Systems
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Tool Type A
Tool Type B
Total
Number Tools
10
5
15
scfm/tool Diversity factor
20
20%
10
40%
Total scfm
200
50
250
Average scfm
40
20
60
Most compressors deliver about 4 scfm per brake hp. Sizing for the average and peak loads
results in very different compressors:
Average load: 60 scfm / 4 scfm/hp = 15 hp
Peak load: 250 scmf / 4 scfm/hp = 60 hp
This example demonstrates why compressors are regularly oversized. Oversizing compressors
typically results in large first costs and large operating costs since many compressors have poor
part-load efficiencies. Partial solutions to the ‘part-load’ dilemma are:



Size AC for average load, but add storage capacity in system for peaks.
Buy multiple smaller compressors so the baseload compressor is generally fully loaded
and the part-load penalty is small for the ‘trim’ compressor. This also adds redundancy
for machine failure and servicing.
Buy a variable-speed compressor or reciprocating compressor with excellent part-load
efficiency.
Inside-Out Approach to Energy Efficient Compressed Air
Application of the whole-system inside-out approach leads to the greatest savings for the least
first cost. First, develop a baseline of the current compressed air system. Next, minimize enduses of compressed air. Next, investigate the distribution system for leaks and excess pressure
drops. Finally, investigate the compressors and dryers for energy saving opportunities.
Minimize End Uses Of Compressed Air
Is compressed air the best source of power for the job?
Use blowers instead of compressed air.
Use valves and sensors to shut of compressed air when not needed
Use Venturi nozzles that amplify flow by up to 20 to 1 to reduce compressed air flow.
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Fix Leaks In The Distribution System
Leaks are expensive! Most compressed air systems lose between 5% and 20% of compressed air
to leaks. We recommend inspecting the compressed-air system for leaks once a week by
listening for leaks when all machinery is off except the air compressor.
Tagged compressed air leak.
To estimate leakage rates from compressed air systems at 100 psig, we use the following table.
Equivalent Hole Diameter Leakage Rate scfm
1/64 "
0.25
1/32 "
0.99
1/16 "
3.96
1/8 "
15.86
1/4 "
63.44
3/8 "
142.74
Source: Compressed Air Systems, DOE/CS/40520-T2, 1984.
The values in this table were computed from the S.A. Moss equation (Ingersoll-Rand Condensed
Air Power Data, 1998)
W(lb/s) = 0.5303 x A (in2) x C x P (psia) / T (R)
where C = 0.97 for a smooth edged hole and C = 0.61 for a sharp edged orifice. The equation can
be modified to show air leakage in standard cubic feet per minute at T = 70 F = 530 R and the
density of air at 70 F is 0.7494 lb/ft3 such that:
V (scfm) = 0.5303 x  / 4 x [D (in)]2 x 0.61 x P (psia) / [ 530( R) x 0.07494 lb/ft3] x 60 s/min
V (scfm) = 8.8356 x [D (in)]2 x P (psia)
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Example:
Calculate the annual electricity cost savings from fixing a leak if a) the compressor is a
reciprocating compressor, and b) the compressor is a screw compressor running in modulation
mode.
Input Data:
Leak: single 1/16-inch diameter
Compressor that uses 0.25 hp per scfm of compressed air
Compressor runs 8,000 hours per year
Electricity costs $0.06 /kWh.
Motor efficiency = 90%
FP0 for reciprocating compressor = 0.0
FP0 for screw compressor with modulation control = 0.70
Calculations:
Unadjusted: 4 scfm x 0.25 hp/scfm x 0.75 kW/hp / 90% x 8,000 hr/yr x $0.06 /kWh = $400 /yr
Adjustment for part-load efficiency: = Unadjusted savings x (1-FP0)
Savings for reciprocating compressor = $400 /yr x (1 - 0.00) = $400 /yr
Savings for screw compressor with modulation control = $400 /yr x (1 - 0.70) = $120 /yr
Use Outside Air For Compression
Theory
W = m cp (T2 - T1)
For polytropic expansion: T2 = T1 (P2/P1)k
W = m cp T1 [(P2/P1)k -1]
Fraction savings = (WT1high - WT1low) / WT1high
Fraction savings = (T1high - T1low) / T1high
Example:
Calculate savings and simple payback from compressing outside air instead of indoor air.
Input Data:
Avoided cost of demand $14.62 /kW-mo Avoided cost of energy = $0.02 /kWh
50-hp compressor running 5,000 hr/yr
Measure: loaded (53 A) 60% of time and unloaded (40 A) 40 % of time at 480 V.
Compress outside air if inside air = 90 F when outside air = 65 F and avg outside air = 50 F
Cost of 16 ft of 3" PVC pipe w/ fg insulation = $50 + (4 hr x $25 /hr labor) = $150
Calculations:
Wloaded = 53 A x 480 V x
3 x 84% PF = 37 kW
Wunloaded = 40 A x 480 V x 3 x 78% PF = 26 kW
Waverage = (60% x 37 kW) + (40% x 26 kW) = 33 kW
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Tout = 65 F, Tin = 90 F, dT = 90 – 65 = 25 F = constant
Toa,avg = 50 F hence T1lo = 50 F, T1high = 75 F
Fraction savings = [(75 + 460) R – (50 + 460) R ] / (75 + 460) R = 4.7%
Wloaded,new = 37 kW x (1-0.047) =35.3 kW
Waverage,new = (60% x 35.3 kW) + (40% x 26 kW) = 31.6 kW
Demand savings = 33 kW – 31.6 kW = 1.4 kW
Demand savings = 1.4 kW x $14.62 /kW-mo x 12 mo/yr = $246 /yr
Usage savings = 1.4 kW x 5,000 hr/yr x $0.02 /kWh = $140 /yr
Total savings: $246 /yr + $140 /yr = $386 /yr
Simple payback = $150 / $386 /yr x 12 mo/yr = 5 months
Reduce Operating Pressure
Theory
W = m cp (T2 - T1)
For polytropic expansion: T2 = T1 (P2/P1)k k = 0.2857 for isentropic expansion of air
W = m cp T1 [(P2/P1)k -1]
Fraction savings = (WPhigh - WPlow) / WPhigh = [(P2high/P1)k - (P2low/P1)k] / [(P2high/P1)k - 1]
Fraction savings = (P2highk - P2lowk) / (P2highk - P1k)
Example:
Calculate savings and simple payback from reducing operating pressure.
Input Data:
Avoided cost of demand $14.62 /kW-mo Avoided cost of energy = $0.02 /kWh
50-hp compressor running 5,000 hr/yr
Measure: loaded (53 A) 60% of time and unloaded (40 A) 40 % of time at 480 V.
Reduce pressure from 110 psig to 100 psig
Calculations:
Loaded power: 53 A x 480 V x
3 x 84% PF = 37 kW
Unloaded power: 40 A x 480 V x 3 x 78% PF = 26 kW
Average power: (60% x 37 kW) + (40% x 26 kW) = 33 kW
Percent full-load power when unloaded: 26 kW / 37 kW = 70%
P1 = 14.7 psia
P2high = 110 psig + 14.7 psi = 124.7 psia
P2low = 100 psig + 14.7 psi = 114.7 psia
Fraction savings: (3.9700 – 3.8763) / (3.9700 – 2.1553) = 5.1%
Loaded power, new: 37 kW x (1-0.051) = 35.1 kW
Average power, new: (60% x 35.1 kW) + (40% x 26 kW) = 31.5 kW
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Demand savings: 33 kW – 31.5 kW = 1.5 kW
Demand savings: 1.5 kW x $14.62 /kW-mo x 12 mo/yr = $263 /yr
Usage savings: 1.5 kW x 5,000 hr/yr x $0.02 /kWh = $150 /yr
Total savings: $263 /yr + $150 /yr = $413 /yr
Implementation cost: none
Simple payback: immediate
Reduce Unloaded Run Time
Rotary screw, centrifugal and some reciprocating compressors run continuously, but only add air
to the compressed air system when the pressure of the system is between the activation
pressures (typically 100 psig to 110 psig). When a compressor is running, but not adding
compressed air to the system, it is said to be unloaded. Unfortunately, rotary screw compressors
typically draw between 50% and 70% of full-load power even when running unloaded. Thus,
there is a huge energy penalty for running unloaded.
The fraction of time a compressor runs unloaded is determined by the relationship between the
capacity of the compressor and the demand for compressed air. If the compressor is under or
properly sized, it will run loaded most of the time. If a compressor is oversized for the load, it will
quickly raise the pressure to the upper bound of the activation pressure and then run unloaded
for an extended period of time.
There are two primary strategies for minimizing the time that compressors run unloaded. The
first is simply to purchase or operate a smaller compressor. In our experience, this is frequently
cost effective whenever a compressor is loaded less than half the time.
Example:
Calculate savings and simple payback for replacing 50-hp screw compressor with 25-hp
reciprocating compressor.
Input Data:
Avoided cost of demand = $14.62 /kW-mo Avoided cost of energy = $0.02 /kWh
50-hp screw compressor generates 4 scfm/hp and runs 2,250 hr/yr
Loaded 10 sec and unloaded 35 sec.
Loaded = 71 Amps PF = 0.84 kW/kVA Voltage = 480 Volts
Unloaded = 45 Amps PF = 0.78 kW/kVA Voltage = 480 Volts
Calculations:
Compressor loaded 10 sec / 45 sec = 25% of the time, thus compressor is oversized.
Loaded power: 71 Amps x 480 Volts x x 0.84 kW/kVA = 49.6 kW
Unloaded power: 45 Amps x 480 Volts x x 0.78 kW/kVA = 29.2 kW
Average power: (25% x 49.6 kW) + (75% x 29.2 kW) = 34.3 kW
Annual elec: 34.3 kW x 2,250 hr/yr = 77,175 kWh/yr
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Percent full-load power when unloaded: 29.2 kW / 49.6 kW = 59%
Power output at full load: 49.6 kW x 90% eff / .75 kW/hp = 59.5 hp
Service Factor: 59.5 hp / 50 hp = 1.19
Average compressed air output: 59.5 hp x 4 scfm/hp x 25% = 59.5 scfm
Percent time loaded for 25-hp recip to supply same output:
59.5 scfm / (25 hp x 4 scfm/hp) = 60%
25-hp recip power: 25 hp x .75 kW/hp / 88.5% = 21.2 kW
25-hp recip elec: 21.2 kW x 60% x 2,250 hr/yr = 28,620 kWh/yr
Demand savings: (34.3 kW – 21.2 kW) x $14.62 / kW-mo x 12 mo/yr = $2,298 /yr
Elec savings: (77,175 kWh/yr – 28,620 kWh/yr) x $0.02 /kWh = $971 /yr
Total savings: $2,298 /yr + $971 /yr = $3,269 /yr
Cost of 25-hp air cooled recip: $7,000
Simple payback = $7,000 / $3,269 /yr x 12 mo/yr = 26 months
The second strategy for reducing the time that compressors run unloaded is to stage multiple
compressors so that unneeded compressors are turned off when not needed, rather than running
unloaded. To stage multiple compressors, set the lower activation pressure of the “baseload”
compressor a few psi higher than the lower activation pressure of the “lag” compressor.
Additional “lag” compressors should activate at increasingly lower pressures. For example, the
“baseload” compressor may be set at 105 psig, the first “lag” compressor at 103 psig and the
second “lag” compressor at 101 psig. If the compressors are staged in this manner, the “lag”
compressors never load unless the “baseload” compressor cannot meet the plant’s demand for
air.
This alone, will not result in energy savings since the lag compressors will continue to run
unloaded while drawing a significant fraction of full-load power. However, most compressors
have a “sleep” or “automatic” mode in which the compressor will turn off if it runs unloaded for 5
or 10 minutes. Staging activation pressures and setting the compressors to run in “sleep” mode
can dramatically reduce energy use while delivering the same system performance. This can be
done manually or using dedicated compressor control equipment, which can also be
programmed to rotate “baseload” duty.
Example:
Calculate savings and simple payback for running 30-hp reciprocating as base compressor instead
of 50-hp screw compressor.
Input Data:
Avoided cost of demand = $14.62 /kW-mo Avoided cost of energy = $0.02 /kWh
Compressors generate 4 scfm/hp and run 6,000 hr/yr
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Current: primary comp is 50-hp screw with 30-hp reciprocating backup.
50-hp cycle time: 5 minutes loaded and 10 minutes unloaded
Fraction time loaded: 5 min / 15 min = 33%
Recommend: baseload 30-hp and run 50-hp in “automatic” mode as lag for peaks
Loaded power = 61 Amps x 480 V x
3 x 84% PF = 42.6 kW
Unloaded power = 44 Amps x 480 V x 3 x 78% PF = 28.5 kW
Average power: (33.3% x 42.6 kW) + (66.6% x 28.5 kW) = 33.2 kW
Annual energy: 33.2 kW x 6,000 hr/yr = 199,200 kWh/yr
Power output at full load: 42.6 kW x 90% eff / 0.75 kW/hp = 51.1 hp
Service Factor: 51.1 hp / 50 hp = 1.02
Average compressed air output: 51.1 hp x 4 scfm/hp x 33% = 67.5 scfm
Percent time loaded for 30-hp recip to supply same output:
67.5 scfm / (30 hp x 4 scfm/hp) = 56%
30-hp recip power: 30 hp x .75 kW/hp / 89% eff = 25.3 kW
30-hp recip elec: 25.3 kW x 56% x 6,000 hr/yr = 85,008 kWh/yr
Demand savings: (33.2 kW – 25.3 kW) x $14.62 / kW-mo x 12 mo/yr = $1,386 /yr
Elec savings: (199,200 kWh/yr – 85,008 kWh/yr) x $0.02 /kWh = $2,284 /yr
Total savings if 50-hp lag never loads: $1,386 /yr + $2,284 /yr = $3,670 /yr
Total savings if 50-hp lag loads and increases demand: $2,284 /yr
Implementation cost: none
Simple payback = immediate
Use Cooling Air For Space Heating
Adiabatic compression of air to 100 psig results in outlet air temperatures of 350 F to 500 F.
When the air is cooled to room temperature, about 80% of the work added to the air is removed
as “waste” heat. In air-cooled compressors, the temperature of the exiting cooling air is typically
between 80 F and 120 F and can be used for space heating or other low-temperature heating
applications. To use this heat for space heating, we recommend equipping air-cooled
compressors with ducts and dampers to direct warm air from compressor into the plant during
winter and out of the plant during summer. This damper system would also keep the compressor
room cool, thereby increasing the lifetime and efficiency of the compressor.
The net amount of heat added to a plant from an air compressor that is currently exhausting the
cooling air depends on the way the air compressor is ventilated. For example, consider the two
scenarios below where warm air from the compressor is being exhausted from the plant.
The first scenario is shown below. Figure 1A is the current ventilation in which outside air is
brought to the compressor then exhausted to the outside during winter. Figures 1B and 1C show
Energy Efficient Compressed Air Systems
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two possible ways of changing the ventilation system to use heat from the compressor to reduce
the space cooling load. In Figure 1B, outside air is still used for cooling, and then directed into the
plant. Figure 1C shows an alternate system with no outside air.
O
O
P
Qc
Qc
Qc
Figure 1A, 1B and 1C. 1A is the current ventilation, 1B is proposed ventilation using outside air,
and 1C is proposed ventilation with no outside air.
If the compressor cooling air is currently coming from the outside and is exhausted to the outside
as in Case 1A, then the net heat from the air compressor to the plant is zero. If the proposed
ventilation system is to continue to draw cooling air from outside and then direct the warm air
into the plant, as in 1B, then the net heat into the plant, Qnet, is
Qnet = Qc – [V x pcp x (Tp – To)]
where pcp is the product of air density and specific heat (0.018 Btu/ft3-F), V is the volume flow
rate of cooling air through the compressor, Qc is the heat from the compressor, Tp is the
temperature of air in the plant ad To is the outside air temperature. The second term in this
expression, V x pcp x (Tp – To), is a penalty for bringing more cold outside air into the plant.
Depending on the outside and plant air temperatures, this penalty could exceed Qc, in which case
the ventilation system would actually be increasing space heating requirements. The preferred
system is shown in 1C. In this case, no additional outside air is brought into the plant and the
Qnet heat gain from the compressor is simply Qc.
Another scenario where warm air from the compressor is being exhausted from the plant is
shown below. In Figure 2A, cooling air is supplied from inside the plant then exhausted to the
outdoors. The recommended system is shown in Figure 2B, where plant air is recirculated
through the compressor.
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O
Qc
P
Qc
Figure 2A and 2B. 2A is the current ventilation system, and 2B is the proposed ventilation system
that recirculates plant air through the compressor.
In Figure 2A, the compressor is actually adding to the space heating load by increasing infiltration
into the plant. The net space heating energy loss is:
Qloss = V x pcp x (Tp – To)
If the ventilation system were changed to B, then the net heating energy gain would be the sum
of the heat added by the compressor, Qc, and the elimination of the previous loss.
Qnet = Qc + [V x pcp x (Tp – To)]
Example:
Calculate savings and simple payback for using heat from air compressor to displace space
heating.
Input Data:
50 hp compressor running 5,000 hr/yr at 480 V
Measure: loaded (53 A) 60% of time and unloaded (40 A) 40 % of time
Initial ventilation as in 1A and proposed as in 1C.
Assume can capture 70% of input energy for space heat for 2,500 hr/yr
Natural gas costs $7 /mmBtu and gas furnace is 80% efficient
Calculations:
Loaded power: 53 A x 480 V x
3 x 84% PF = 37 kW
Unloaded power: 40 A x 480 V x 3 x 78% PF = 26 kW
Average power: (60% x 37 kW) + (40% x 26 kW) = 33 kW
Space heat: 33 kW x 70% x 2,500 hr/yr x 3,413 Btu/kWh = 197 mmBtu/yr
Gas cost savings: 197 mmBtu/yr / 80% x $7 /mmBtu = $1,724 /yr
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Implementation cost for ducts and dampers: $1,000
Simple payback: $1,000 / $1,724 /yr x 12 mo/yr = 7 months
References
1) Compressed Air Systems, U.S. Dept. of Energy, DOE/CS/40520-T2, 1984.
2) Cengal, Y. and Boles, M., Thermodynamics, 1998, WGB-McGraw-Hill.
3) Condensed Air Power Data, Ingersoll-Rand, 1984.
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