Components
Accumulators ... 5
Actuators
linear ... 54
rotary ... 152
Air
compressors ... 11
dryers ... 15
filters... 19
logic ...21
motors ... 26
regulators ... 30
Airline lubricators ...25
Boots & bellows...32
Carriers for hose & cable ...34
Cartridge valves ... 35
Clamps for pipe, hose, tubing...37
Control networks... 39
Cylinders ... 54
Directional-control valves... 62
Electrohydraulic
proportional valves ... 68
servovalves ... 68
Fittings ...75
Flow-control valves... 78
Flow meters...143
Flushing procedures ...84
Heat exchangers ... 87
Hydraulic
filters... 91
fluids ... 103
hose... 110
manifolds ... 136
pumps ... 119
Pressure-control circuits . . . 215
Pump-unloading circuits . . . 217
Regenerative circuits . . . 219
Safety circuits . . . 220
Sequencing circuits . . . 221
Series motor circuit . . . 210
Speed-control circuits
Hydraulic . . . 222
Pneumatic . . . 212
Synchronizing circuits . . . . 225
Fluid power graphic symbols
Accumulators . . . 242
Air motors . . . 230
Compressors . . . 229
Directional-control valves . 234
Cylinders . . . 231
Filters . . . 242
Flow-control valves . . . 239
Hydraulic motors . . . 230
Intensifiers . . . 233
Proportional valves . . . 236
Pressure-control valves . . . . 237
Pumps . . . 229
Sources of energy . . . 24
Logic symbols
. . . 247
Glossaries
Compressed air terms ... 12
Fluid power terms ... 249
Hydraulic fluids... 108
Organizations with interest
in fluid power ...
255
General index
...
256
Hydrostatic transmissions ... 128
Intensifiers... 133
Motors
air... 26
hydraulic ... 114
low-speed/high-torque... 134
Pressure-control valves ... 138
Pressure gages ... 143
Pressure switches ... 146
Reservoirs ... 148
Rotary actuators ...152
Seals & packings ... 154
Shock absorbers ... 163
Transducer technology ... 167
Vacuum technology... 191
Basic circuits
Index of basic circuits ... 196
Accumulator circuits . . . 197
Air lubrication circuits . . . . 199
Cylinder locking circuits . . 200
Deceleration circuits . . . 202
Decompression circuits . . . 204
Electrical control circuits . . 206
Filter circuits . . . 208
Hydraulic filter circuits . . . . 208
Hydraulic motor circuits . . . 209
Hydraulic speed control . . . 222
Intensifier circuits . . . 211
Locking circuits . . . 200
Meter-in circuits . . . 222
Meter-out circuits . . . 222
Motor braking circuit . . . 209
Parallel motor circuit . . . 210
Pneumatic speed-control . . 212
Engineering
A
ccumulators usually are in-stalled in hydraulic systems to store energy and to smooth out pulsations. Typically, a hydraulic sys-tem with an accumulator can use a smaller pump because the accumulator stores energy from the pump during pe-riods of low demand. This energy is available for instantaneous use, re-leased upon demand at a rate many times greater than what could be sup-plied by the pump alone.Accumulators also can act as surge or pulsation absorbers, much as an air dome is used on pulsating piston or ro-tary pumps. Accumulators will cushion hydraulic hammer, reducing shocks caused by rapid operation or sudden starting and stopping of power cylin-ders in a hydraulic circuit.
There are four principal types of ac-cumulators: the weight-loaded piston type, diaphragm (or bladder) type, spring type, and the hydro-pneumatic piston type. The weight-loaded type was the first used, but is much larger and heavier for its capacity than the modern piston and bladder types. Both weighted and spring types are infre-quently found today. Hydro-pneumatic accumulators, Figure 1, are the type most commonly used in industry.
Functions
Energy storage —
Hydro-pneu-matic accumulators incorporate a gas in conjunction with a hydraulic fluid. The fluid has little dynamic power-storage qualities; typical hydraulic fluids can be reduced in volume by only about 1.7% under a pressure of 5000 psi. (However, this relative
incompressibil-fluid. The piston in turn, forces the fluid from the cylinder into the system and to the location where useful work will be accomplished.
Pulsation absorption — Pumps, of
course, generate the required power to be used or stored in a hydraulic system. Many pumps deliver this power in a pulsating flow. The piston pump, com-monly used for its high pressure
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ity makes them ideal for power trans-mission, providing quick response to power demand.) Therefore, when only 2% of the total contained volume is re-leased, the pressure of the remaining oil in the system drops to zero.
On the other hand, gas, the partner to the hydraulic fluid in the accumulator, can be compressed into small volumes at high pressures. Potential energy is stored in the compressed gas to be re-leased upon demand. Such energy can be compared to that of a raised pile driver ready to transfer its tremendous energy upon the pile. In the piston type accumulator, the energy in the com-pressed gas exerts pressure against the piston separating the gas and hydraulic
Shell Charging valve Bladder Poppet Spring Hydraulic cap Charging valve Gas cap Body Piston Hydraluic cap
Bladder Gas Piston
Fig. 1. Cross-sectional views of typical of bladder and piston-type accumulators.
For more information about using accumulators in hydraulic sys-tems, please refer to the Basic Cir-cuits section which appears else-where in this handbook.
bility, can produce pulsations detri-mental to a high-pressure system. An accumulator properly located in the system will substantially cushion these pressure variations.
Shock cushioning — In many fluid
power applications, the driven member of the hydraulic system stops suddenly, creating a pressure wave that travels back through the system. This shock wave can develop peak pressures sev-eral times greater than normal working pressures and can be the source of sys-tem failure or objectionable noise. The gas cushion in an accumulator, prop-erly placed in the system, will mini-mize this shock.
An example of this application is the absorption of shock caused by sud-denly stopping the loading bucket on a hydraulic front end loader. Without an accumulator, the bucket, weighing over 2 tons, can completely lift the rear
wheels of a loader off the ground. The severe shock to the tractor frame and axle, as well as operator wear and tear, is overcome by adding an adequate ac-cumulator to the hydraulic system.
Supplementing pump flow — An
accumulator, capable of storing power can supplement the hydraulic pump in delivering power to the system. The pump stores potential energy in the ac-cumulator during idle periods of the work cycle. The accumulator transfers this reserve power back to the system when the cycle requires emergency or peak power. This enables a system to utilize a much smaller pump, resulting in savings in cost and power.
Maintaining pressure — Pressure
changes occur in a hydraulic system when the liquid is subjected to rising or falling temperatures. Also, there may be pressure drop due to leakage of hy-draulic fluid. An accumulator
compen-sates for such pressure changes by de-livering or receiving a small amount of hydraulic fluid. If the main power source should fail or be stopped, the ac-cumulator would act as an auxiliary power source, maintaining pressure in the system.
Fluid dispensing — An
accumula-tor may be used to dispense small vol-umes of fluids, such as lubricating greases and oils, on command.
Operation
When sized and precharged prop-erly, accumulators normally cycle be-tween stages (d) and (f), Figure 2. The piston will not contact either cap in a piston accumulator, and the bladder will not contact the poppet or be com-pressed so that it becomes destructively folded into the top of its body.
Manufacturers specify recom-mended precharge pressure for their ac-Gas (a) (b) (d) (f) (e) (c)
Fig. 2. Six stages of operation for bladder- and piston-type accumulators: stage (a), the accumulator is empty -- no gas charge; stage (b), the accumulator has been precharged with dry nitrogen; stage (c), system pressure exceeds precharge pressure, and hydraulic fluid flows into accumulator; stage (d), system pressure peaks, maximum fluid has entered the accumulator, and system relief opens; stage (e), system pressure drops, precharge pressure forces fluid from the accumulator and into the system; and stage (f), system pressure reaches minimum needed to do work.
cumulators. In energy-storage applica-tions, a bladder accumulator typically is precharged to 80% of minimum hy-draulic system pressure and a piston ac-cumulator to 100 psi below minimum system pressure. Precharge pressure determines how much fluid will remain in the accumulator at minimum system pressure.
Correct precharge involves accu-rately filling an accumulator’s gas side with a dry inert gas, such as nitrogen, while no hydraulic fluid is in the fluid side. Accumulator charging then be-gins when hydraulic fluid is admitted into the fluid side, and occurs only at a pressure greater than the precharge pressure. During charging, the gas is compressed to store energy.
A correct precharge pressure is the most important factor in prolonging ac-cumulator life. The care with which precharging must be accomplished and maintained is an important considera-tion when choosing the type of accumu-lator for an application, all else being equal. If the user tends to be careless about gas pressure and relief valve set-tings, or adjusts system pressures with-out making corresponding adjustments to precharge pressure, service life may be shortened, even if the correct type of accumulator was selected. If the wrong accumulator was selected, premature failure is almost certain.
Mounting position
The optimum mounting position for any accumulator is vertical with the hy-draulic port down. Piston models can be horizontal if the fluid is kept clean. When solid contaminants are present or expected in significant amounts, hori-zontal mounting can result in uneven or accelerated seal wear. Maximum ser-vice life can be achieved in the
horizon-tal position with multiple piston seals to balance the piston’s parallel surface. A bladder accumulator also can be mounted horizontally, Figure 3, but un-even wear on the bladder as it rubs against the shell while floating on the fluid can shorten life. The amount of damage depends on fluid cleanliness, cycle rate, and compression ratio (de-fined as maximum-system-pressure/ minimum-system-pressure). In ex-treme cases, fluid can be trapped away from the hydraulic end, which reduces output or may elongate the bladder to force the poppet closed prematurely.
Sizes and outputs
Available sizes and capacities also influence which accumulator type to choose. Piston accumulators of a par-ticular capacity often are supplied in a choice of diameters and lengths, Table 1. Furthermore, piston designs can be built to custom lengths for little or no price premium. Bladder accumulators are offered only in one size per capac-ity, with fewer capacities available.
The inherently higher output of the piston accumulator may make it the best alternative when space is tight.
Table 1 lists outputs for 10-gal piston and bladder accumulators operating isothermally as auxiliary power sources over a range of minimum sys-tem pressures. The differences in precharge pressure, columns 3 and 4, (determined by 80% of minimum sys-tem pressure for bladder models, 100 psi below minimum for piston) lead to a substantial difference in outputs, columns 5 and 6.
To prevent excessive bladder defor-mation and high bladder temperatures, also note in Table 1 that bladder accu-mulators should be specified with com-pression ratios greater than 3:1.
Multiple components
Although bladder designs are not available in sizes over 40 gal, piston designs are currently supplied up to 200 gal in a single vessel. Economics and available installation space have led engineers to consider multiple com-ponent installations. Two of these can cover most high-output applications.
The installation in Figure 4 consists of several gas bottles serving a single piston accumulator through a gas mani-fold. The accumulator portion must be
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Fig. 3. Horizontally mounted accumulator can cause uneven bladder wear and trap fluid away from the hydraulic valve.
Gas manifold Fluid
Fig. 4. Piston accumulators used in conjunction with gas bottles. Table 1–Relative outputs, 10 gal accumulator
System pressure, psi
maximum 1 minimum 2 bladder 3 piston 4 bladder 5 piston 6
1.5 3000 2000 1600 1900 2.53 3.00 2.0 3000 1500 1200 1400 3.80 4.41 3.0 3000 1000 800 900 5.06 5.70 6.0 3000 500 – 400 – 6.33 Compression ratio 1/2 Recommended precharge, psi Output, gal
A C C U M U L A T O R S
sized so the piston does not repeatedly strike the caps while cycling. One draw-back of this arrangement is that a single seal failure could drain the gas system. Because gas bottles often are less
ex-pensive than accumulators, one advan-tage of this setup might be lower cost.
Several accumulators, either piston or bladder design, can be mounted on a hydraulic manifold, Figure 5. If using piston accumulators, the piston with the least friction will move first and oc-casionally could bottom on the hy-draulic cap. In slow or infrequently used systems, this is insignificant.
Gas bottle installations
Remote gas storage offers flexibility in large and small systems, Figure 6. The gas bottle concept is generally de-scribed with this simple formula: accu-mulator size minus required fluid out-put equals gas bottle size. For example, an application that calls for a 30-gal ac-cumulator may only require 8 to 10 gal of fluid output. This application, there-fore, could be satisfied with a 10-gal accumulator and a 20-gal gas bottle.
An accumulator used with remote gas storage generally has the same size port at the gas end as at the hydraulic end to allow unimpeded flow of gas to and from the gas bottle. The gas bottle has an equivalent port in one end and a gas charging valve at the other. These two-piece accumulators can be configured or bent at any angle to fit available space.
The gas bottle concept is suitable for either bladder or piston accumulators. Note that bladder accumulators require a special device called a transfer
bar-rier at the gas end to prevent extrusion
of the bladder into the gas bottle piping. Again, a piston accumulator should be sized to prevent piston bottoming at either end of the cycle. Bladder designs should be sized to prevent filling to more than 85% or discharging to more than 85% empty. The flow rate be-tween the bladder transfer barrier and its gas bottle will be restricted by the
neck of the transfer barrier tube. Be-cause of these drawbacks, bottle/blad-der accumulators should be reserved for special applications.
Flow rates and response times
Table 2 suggests maximum flow rates for representative accumulator sizes and types. The larger standard bladder designs are limited to 220 gpm, although the rate can be boosted to 600 gpm using an extra-cost, high-flow port. The poppet controls flow rate; ex-cessive flow causes the poppet to close prematurely. Multiple accumulators mounted on a common manifold are needed to achieve flows that are greater than 600 gpm.
Allowable flow rates for piston accu-mulators generally exceed those for bladder designs. Flow is limited by pis-ton velocity, which should not exceed 10 ft/sec to avoid piston seal damage. In high-speed applications, high seal contact temperatures and rapid decom-pression of nitrogen that has permeated into the seal material can cause blisters, cracks, and pits in the rubber.
Bladder accumulators respond more quickly to system pressure variations than do piston types for two reasons:
1. Rubber bladders do not have to
overcome the static friction which a piston seal must, and
2. The piston mass does not need to
be accelerated and decelerated. In practice, though, the difference in response may not be as great as com-monly believed, and is probably insig-nificant in most applications.
Shock suppression
Tests at the University of Wisconsin, Madison, indicate that shock control does not necessarily demand a bladder accumulator. With system flow at a
Table 2–Maximum recommended accumulator flow rates
Bladder capacity 2 1 qt 100 60 – 4 1 gal 400 150 – 6 2.5 gal 800 220 600 7 1200 220 600
9 larger than 2.5 gal 2000 220 600
12 3400 220 600
Piston
bore, in. Piston Standard High-flow
Tubing length makes pump-to-servovalve distance 118 ft LVDT
Fig. 6. A small accumulator may do the job if it is remotely connected to an auxiliary gas bottle.
Fig. 7. Test circuit to generate and mea-sure shock waves in system.
A C C U M U L A T O R S
Bladder Gpm at 3000 psi Fig. 5. Several accumulators may be
man-ifolded to provide large system flows.
Fluid manifold
nominal 30 gpm in the test circuit, Fig-ure 7, an internally piloted directional control valve, 118 ft away from the pump, closes to generate a shock. As the shock wave travels from the valve back through the hydraulic lines and around corners and various restrictions, some portion of its energy is consumed while accelerating the mass of fluid in the lines.
With 11/4-in. tubing, a 2750-psi relief valve setting, and no accumulator in the circuit, oscilloscope trace A, Figure 8, shows a pressure spike of 385 psi over the relief valve setting. Adding a 1-gal piston accumulator at the valve reduces the transient to 100 psi over relief valve setting, trace B. Substituting a 1-gal bladder accumulator cuts the transient to 78 psi over relief valve setting, trace
C, only 22 psi better than the
piston-type protection.
A second, similar test with 5/8-in. tubing and a relief valve setting of 2650 psi results in a pressure spike of 2011 psi over relief valve setting without an accumulator, trace A, Figure 9. A pis-ton accumulator damps the transient to 107 psi over relief valve setting, trace
B, while a bladder accumulator damps
the transient to 87 psi over relief valve setting, trace C. The difference be-tween accumulator types in shock sup-pression again was negligible.
Servo equipment
Another common misconception says that all servo applications require a bladder accumulator. Experience shows that only a small percentage of servos require response times of 25 ms or less, the region where the difference in response between piston and bladder accumulators becomes material. Blad-der accumulators should be used for applications requiring less than a 25-ms response, and either type when re-sponse of 25 ms or greater is adequate.
Setup and maintenance: precharging
On newly repaired bladder accumu-lators, the shell ID should be lubricated with system fluid before precharging. This fluid acts as a cushion, and lubri-cates and protects the bladder as it un-winds and unfurls. When precharging begins, the initial 50 psi of nitrogen should be introduced slowly.
Neglecting these precautions could result in immediate bladder failure.
High-pressure nitrogen, expanding rapidly and thus cold, could channel the length of the folded bladder and con-centrate at the bottom. The chilled brit-tle rubber expanding rapidly could rup-ture in a starburst pattern, Figure 10(a). The bladder also could be forced under the poppet, resulting in a C-shaped cut in the bladder bottom, Figure 10(b).
The fluid side of piston accumulators should be empty during precharging so that gas-side volume is at a maximum. Little damage, if any, can take place during precharging.
Too high a precharge pressure or re-ducing the minimum system pressure without a corresponding reduction in precharge pressure may cause operat-ing problems or damage to accumula-tors. With excessive precharge pres-sure, a piston accumulator will cycle between stages (e) and (b), Figure 2, and the piston will range too close to the hydraulic end cap. The piston could bottom at minimum system pressure to reduce output and eventually cause damage to the piston and its seal. The bottoming of the piston often can be heard; the sound serves as a warning of impending problems.
Too high a precharge in a bladder ac-cumulator can drive the bladder into the poppet assembly when cycling between stages (e) and (b), Figure 2. This could cause fatigue failure of the spring and poppet assembly, or a pinched and cut bladder if the bag gets trapped beneath the poppet as it is forced closed. Too high a precharge pressure is the most common cause of bladder failure.
Fig. 10. Starburst rupture in end of bladder, (a), could indicate loss of elasticity of blad-der material due to embrittlement from cold nitrogen gas during precharge. If blad-der is forced unblad-der poppet, (b), bladblad-der could sustain C-shaped cut from poppet.
Fig. 8. Graph indicates results of shock wave tests. Fig. 9. Results of second test using smaller-diameter tubing.
Trace C Trace B Trace A Time – ms 0 1000 2000 2000 4000 Pressure – psi 4000 2000 1000 2000 0 Time—ms Pressure—psi Trace A Trace B Trace C (a)
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A C C U M U L A T O R S
Too low a precharge pressure or an increase in system pressure without a compensating increase in precharge pressure also can cause operating problems, with possible accumulator damage. With no precharge in a pis-ton accumulator, the pispis-ton likely will be driven into the gas end cap and probably will remain there. A single contact is unlikely to cause damage.
For bladder accumulators, too low or no precharge can have severe con-se qu en ce s. T he b l a dde r m ay b e crushed into the top of the shell, then may extrude into the gas valve and be punctured, Figure 11. One such cycle is sufficient to destroy a bladder.
Pis-ton accumulators, therefore, are more tolerant of improper precharging.
External forces
Any application subjecting an accu-mulator to acceleration, deceleration, or centrifugal force may have a detri-mental effect on accumulator opera-tion. Forces along the axis of an accu-mulator’s tube or shell normally have little effect on a bladder model but may increase or decrease gas pressure in a piston type because of the mass of the piston affects the force.
Forces perpendicular to an accumu-lator’s axis should not affect a piston model, but fluid in a bladder
accumula-tor may be thrown to one side of the shell, displacing the bladder and flat-tening and lengthening it, Figure 12. With this distortion, fluid discharge could cause the poppet valve to pinch and cut the bladder.
Failure prediction
Several methods can be used to mon-itor the precharge pressure of piston ac-cumulators:
With the hydraulic system shut down — A pressure transducer or gage
located in the gas end cap, Figure 13(a), indicates the true precharge pressure after a working hydraulic system has cooled, and the accumulator does not contain fluid.
With the hydraulic system operat-ing — On request, accumulator
manu-facturers will install a piston-position sensor in an accumulator’s hydraulic end cap, Figure 13(b). This sensor can be connected to a number of electronics packages. With an accurate precharge and after enough system operation for thermal stability, the electronics can be calibrated to provide continuous read-out of precharge pressure that corre-sponds accurately to the true precharge.
With the accumulator coupled to a gas bottle — A ferrous or nonferrous
sensor can be installed in the accumula-tor gas end cap, Figure 13(c), to detect when the piston comes within 0.125 in. of the cap. This warning indicates that precharge pressure has dropped, and the system should be shut down and checked.
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Pressure transducer Fluid (a) Sonar absolute position sensor Pressure transducer Fluid (b) Pressure transducer Fluid (c) Hall proximity sensorFig. 11. If pressure fluid is allowed into uncharged accumulator, bladder could be crushed or extruded into gas valve and punctured.
Fig. 12. Forces applied perpendicular to bladder accumulator vertical axis can dis-tort bladder shape and risk bladder punc-ture.
Fig. 13. With pressure transducer mounted in cap of piston-type ac-cumulator, (a), actual precharge will be indicated after working system has become dormant and cooled. Piston-position sensor, (b), can provide continuous read-out of precharge when connected to proper electronics package. With Hall-effect sensor installed, (c), close proximity of piston to end cap can be indicated. Note: pistons in (b) and (c) must be flat for use with sonar or Hall-effect sensors. Acceleration
A C C U M U L A T O R S
C
ompressed air has become one of the most important power media used in industry. What kind of equipment makes the compres-sion process tick? Two basic types of machines compress ambient air for in-dustrial use: positive-displacement and dynamic air compressors.In positive-displacement compres-sors, ambient air is isolated in a volume that subsequently is mechanically re-duced to increase the air’s pressure. The action may use a crankshaft and re-ciprocating pistons — much like the fa-miliar internal combustion engine — or rotary elements. The most common ro-tary elements are sliding vanes that move radially, and male and female ro-tors that mesh as they turn.
In dynamic compressors, the me-chanical action of rotating impellers ac-celerates ambient air as it passes through the machine. The additional ki-netic energy is converted into pressure energy downstream. Dynamic com-pressors are identified as centrifugal or axial — depending on the manner in which air flows through them.
Thermodynamics
The various compression processes are based on the ideal gas laws of ther-modynamics. Neither air nor other gases meet all the assumptions implied in these perfect gas laws. However, some knowledge of these laws can be com-bined with information gained from ex-perimentation to permit an engineering analysis of the compression process.
Compression efficiency in any com-pressor is compared with two theoreti-cal standards —isothermal and
adia-batic. (Neither type occurs in an actual
compressor because of the unavoidable losses of the real world.) Isothermal compression would occur if the air tem-perature were kept constant as pressure increases. To keep temperature con-stant, the heat of compression would
facturer provides the driver and com-pressor as an assembly according to these general technical considerations:
● voltage and frequency requirements ● any current restrictions —
particu-larly kVA inrush during starting
● motor-to-compressor speed match ● torque requirements for starting and
running
● power factor considerations ● cycling considerations ● proper protective enclosure ● ambient temperature range ● desired efficiency, and ● anticipated service factor.
Economic considerations involving the drive motor are motor cost, mainte-nance cost, operating cost, and — most important — an analysis of accessory equipment needed to operate the motor. Along with the starter and motor con-trols, these include any special trans-former requirements and cost of addi-tional power lines.
Reciprocating compressors
The original industrial air compres-sors were reciprocating machines and they still are offered in more models and sizes than perhaps any other type. Horsepower ratings today range from fractional to around 600; available pres-sures can be 6000 psi and higher. This variety makes it possible to find a recip-rocating compressor small enough to operate a single function on a machine or large enough to supply a small manu-facturing facility. A number of differ-ent physical configurations, Figure 1, may even make it possible to match a compressor to the particular space where it must be installed. Reciproca-tors are the most efficient compressors for the majority of applications. They can be fitted with control systems which match their output almost ex-actly to operating demands. Modern electronic pressure sensors join sophis-ticated computer-control systems to
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have to be removed continually. This perfect cooling cannot be accomplished in actual practice. The isothermal equa-tion is a statement of Boyle’s Law:
P1V1= P2V2
Adiabatic compression would occur if no heat were transferred to or from the outside during the process. True adi-abatic compression also is not attained in practice. The adiabatic equation is:
P1V1 k
= P2V2 k
where k is the ratio of the specific heat at constant pressure to the specific heat at constant volume, cp/cv. For dry air, k ≈1.4.
Dynamic compressors follow a poly-tropic compression cycle in which the relationship of pressure and volume is held constant. The polytropic equation is:
P1V1 n
= P2V2 n
where n is a constant, determined ex-perimentally for the particular type of machine involved.
Significant power can be saved in any type of compressor by using the multi-stage principle. In these com-pressors, the output from the first stage is fed to the inlet of the second stage, and so on. Cooling between stages saves more power, because the com-pression process — adiabatic or poly-tropic — is non-linear with an expo-nential constant greater than unity.
Drivers
Electric motors are the most widely used compressor drivers for industrial facilities. (Steam and natural gas en-gines are other possibilities with limited applications.) Internal-combustion en-gines serve portable compressors. In al-most every case, the compressor
manu-Some basic information
about the machines that
produce industrial
compressed air.
Air
quence a gang of different-size air com-pressors for maximum efficiency.
Reciprocating compressors have one or more cylinders, each fitted with a piston driven by a crankshaft through a connecting rod. Each cylinder also has intake and discharge valves, and some means for cooling the mechanical parts. Ambient air is drawn into the cyl-inder during its suction stroke. At the end of the suction stroke, the crankshaft reverses the piston’s direction and the air is compressed and expelled during the discharge stroke.
When only one end of the piston
contacts the air, the compressor is iden-tified as single-acting. when both ends of the piston act on the air, the com-pressor is double-acting. Obviously, a double-acting compressor discharges approximately twice as much air per cylinder per cycle as a comparable-size single-acting machine.
As in other high-cycling machines, lubrication and cooling are important to the operation of reciprocating compres-sors. Depending on compressor size, splash lubrication, pressurized crank-cases, or pumped lubrication may pro-vide the former of these functions.
Wa-ter is the most-common coolant for air-compressor cylinders, intercoolers, and aftercoolers, although some smaller models may be air-cooled.
Rotary-vane compressors
The basic elements of this type of compressor are a cylindrical case with an internal eccentrically mounted rotor, Figure 2. Vanes fit into radial slots in the rotor, and move centrifugally out-ward until they meet the case’s inside surface as the rotor turns. When the vanes pass inlet ports in the case, they form pockets to trap air. These pockets decrease in size due to the rotor’s ec-centric location. The trapped air is compressed, and then expelled when the pocket reaches a discharge port.
Because of sliding friction in the slots and tip wear against the case, the vanes are the most wear-prone parts of this type compressor. Vane length, vane rubbing speed, and bending forces on the extended vanes all tend to limit
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A I R C O M P R E S S O R S
Fig. 1. Reciprocating compressors. Upper image: Kaeser Compressor’s oil-less Airbox, in sound-damp-ening enclosure with anti-vibration frame and pad-ding, generates noise levels only 66 to 69 dB(A). Models from 4 to 15 hp have TEFC motors. Lower
im-age: Thomas Industries’ Model HP-1500 11⁄2-hp piston
air compressor can flow 5.2 cfm at 100 psi.
Adiabatic compression occurs when
no heat is transferred to or from the air during compression.
Aftercooling — the cooling of air
af-ter it has been compressed to lower its temperature and precipitate con-densed vapors.
Boyle’s Law states that the absolute
pressure of a fixed mass of gas varies inversely as the volume, provided the temperature remains constant.
Charles’ Law states that the volume
of a fixed mass of gas varies directly with the absolute temperature, pro-vided that pressure remains constant.
Clearance is the volume of a
recipro-cating compressor’s cylinder not swept by the piston’s movement. It includes space between the piston and the head at the end of the compression stroke, and typically is expressed as a percentage of cylinder displacement. The clearance may be different at the two ends of a double-acting cylinder.
Compression efficiency is the ratio of
the theoretical work required (in a given process) to the actual work re-quired to be done to compress and de-liver the air. Expressed as a percent-age, compression efficiency accounts for fluid-friction losses, leakage, and thermodynamic variations from the theoretical process.
Compression ratio is the ratio of the
absolute discharge pressure to the ab-solute intake pressure.
Density is the weight of a given
vol-ume of air, usually expressed in lb/ft3
at standard temperature and pressure.
Displacement is the net volume
swept by the moving parts of a com-pressor per unit of time. This term ap-plies only to positive-displacement compressors.
Free air is air at the atmospheric
con-ditions at a specific location.
Ideal gases follow the perfect gas
laws without deviation. There are no real ideal gases, but they provide a common starting point for calcula-tions and correccalcula-tions.
Intercooling is the process of cooling
air between stages of compression to liquefy condensed vapors and save power by reducing the temperature of air entering the next stage.
Isothermal compression occurs
when the temperature of the air re-mains constant during compression.
Mechanical efficiency is the ratio of
the indicated horsepower to the actual shaft horsepower.
Polytropic compression occurs
when heat is transferred to or from the air at a precise rate during compres-sion so that PVn
is constant.
Pumping or surge is the reversal of
flow within a dynamic compressor. It takes place when insufficient pressure is generated to maintain flow.
Standard pressure and tempera-ture is generally defined as 68º F and
14.70 psia with a relative humidity of 36%.
Volumetric efficiency is the ratio of
the actual volume of air admitted (at a specified temperature and pressure) to the full piston-displacement vol-ume — obviously for reciprocating compressors only.
tended periods of time with little main-tenance required, particularly when matched with the sophisticated control and warning systems available today. These machines are built in two ver-sions: wet (or oil-flooded) and dry. In either one, air is compressed by the ac-tion of two intermeshing rotors which turn inside a housing. Filtered ambient air is drawn into the voids formed as male and female rotors unmesh, then trapped and sealed as these voids pass the intake ports. As rotation continues, the volume of the voids decreases and pressure increases as the rotors remesh. In oil-flooded machines, cooling oil is sprayed into the housing during com-pression. This oil absorbs the heat of compression as it is generated, lubri-cates all dynamic contact surfaces, and also forms a seal between rotors and be-tween rotors and housing walls. The oil then is separated from the compressed air, cooled, filtered, and returned to the injection point for re-use.
Dry screw compressors are ma-chined to extremely close tolerances so that the two rotors do not touch. There-fore lubrication of the compression chamber is unnecessary. These com-pressors can operate at high rotational speeds. A cooling system still is needed to remove the heat of compression.
Centrifugal compressors
Centrifugal compressors develop pressure within themselves,
indepen-dent of load — but the load determines the flow to be handled. This general statement is, of course, limited by the physical size of the machine and the power of its driver.
In its simplest form, a centrifugal compressor is a stage, single-flow machine with its impeller over-hung on its drive, Figure 4. Air enters the unit through the inlet nozzle, which is proportioned so that the air arrives at the impeller with a minimum of shock and turbulence. The impeller receives air from the inlet nozzle and dynami-cally compresses it. The impeller also sets the air in motion, achieving a ve-locity somewhat less than the tip speed of the impeller.
A diffuser chamber surrounds the impeller and receives air leaving the impeller. The diffuser serves to gradu-ally reduce the velocity of the air and convert its velocity energy to a higher pressure level. A volute casing sur-rounds the diffuser and repeats the pro-cedure, collecting the air, reducing its velocity further, and recovering addi-tional velocity energy.
The stresses that are permissible in the impeller limit the maximum dis-charge pressure that may be obtained from this single-stage unit. For higher pressures, the answer again is multiple staging— with multiple impellers and passages to take air from each diffuser
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A I R C O M P R E S S O R S
Fig. 2. A typical rotary-vane compressor has oil injected during compression to absorb some heat of compression. Air exiting from vane and screw compressors is delivered to a separator where liquid oil is removed.
Fig. 3. TS Series tandem air compressor — available in 100- to 600-hp sizes — has two sets of rotors that di-vide compression process equally between two stages to develop flow capacities from 515 through 3100 cfm.
the flow capacity and pressure ratings of rotary-vane compressors.
Oil frequently is injected into rotary-vane compressors to help with sealing, provide lubrication, and absorb some of the heat of compression. In older in-stallations, the combination of high heat and oil sometimes formed sludge and varnish deposits. Modern lubricat-ing oils have eliminated that situation.
Rotary-screw compressors
Rotary-screw compressors, Figure 3, have become the dominant design be-cause they literally offer more horse-power per dollar. They can run for
ex-Fig. 4. Cutaway view of single-stage, single-inlet centrifugal compressor with closed-type impeller. Electric drive motor is at left center.
to the inlet of the succeeding stage. Axial compressors accelerate the air in a direction generally parallel to the rotating shaft. Each pair of rotating and stationary blades form a stage, but pres-sure rise per stage is small, so the usual axial compressor must have multiple stages to produce typical shop air.
While centrifugal machines deliver practically constant pressure over a considerable range of capacities, axial compressors have a substantially con-stant flow delivery at variable pres-sures. Note that these characteristics also mean the flow from a centrifugal compressor must be greatly reduced to increase the pressure ratio, while an ax-ial compressor can develop a substan-tial increase in pressure with a modest reduction in flow rate. In general, cen-trifugal machines have a wider stable range than axial compressors. Because of their more-or-less straight-line flow, axial machines tend to be smaller in di-ameter than centrifugal machines, and
are apt to be longer. Their efficiency also tends to be slightly higher.
Both types of dynamic compressors are subject to a stable maximum flow which is limited by an inlet choking phenomena. They also exhibit non-sta-ble minimum flow which is affected by an outlet surge phenomena.
Flow rates and pressure
Regardless of basic design, the role of any industrial air compressor is to produce a rate of flow at a level of pres-sure satisfactory to meet the demands of the facility in which it will operate. For many industrial plants, this is a dy-namic situation, with fluctuating loads every day and a high probability of in-creased demand — due to new equip-ment and operations — in the future. Obviously, an informed guess about fu-ture compressed-air requirements will be valuable when considering a new or augmented compressor installation.
Assuming that most industrial facili-ties want compressed air in a pressure range from 100 to 125 psi, a good rule of thumb is that central, multi-stage compressors in the 100-hp size range will deliver about 5 scfm/hp. Delivery improves slightly as compressor size and horsepower rating increase, and conversely,decreases for smaller units.
Another aspect of pressure: if only one section of a facility requires rela-tively high air pressure, consider a sep-arate, high-pressure compressor to sup-ply that area. Or install a booster compressor to increase shop pressure to the higher level in that section.
Compressor selection
Those who purchase plant air pressors today generally buy a com-pletely packaged unit augmented by many other components besides the ba-sic compressor. Evaluating these sup-porting components and their ability to help the compressor match output to operating conditions is, perhaps, more important than the choice of the type of compressor. When selecting a com-pressor package for a particular appli-cation, users should always consider every facet of the package, including:
● compressor capacity — pressure and
flow ratings, including performance at the low or high ends of the operating pressure range, and at the altitude where it will be installed
● provision for air- or water-cooling ● electric motor drive
● electric motor starter ● type of control ● serviceability ● maintainability
● type of enclosure and its function ● required auxiliary equipment ● operating noise level
● special accessories, such as unit or
compressor prefilter; outdoor enclo-sure; cold or hot weather package; heat-recovery package; aftercooling, reheating, and filtration of discharge air; even skid or wheel mounting for portable units; and, of course
● cost and warranty.
A word of caution on capacity rat-ings: be sure the compressor manufac-turer provides data about the actual volume of air the package will deliver under your operating conditions. Some manufacturers rate their machines in
delivered air, others in flange-to-flange
or bare compressor performance. To compare bare-compressor ratings, they must be reduced to account for inlet fil-tration, cooling, and other losses. Such losses can range as high as 5 to 7%.
Power cost
At one time, many users of industrial compressed air considered it free — or at least so inexpensive that its cost was negligible. Steeply rising energy costs over the past twenty years have put that misconception to rest. Power cost is the largest single item in the operating cost of an air compressor. Over the typical long life of the machine, power costs will be large multiples of the initial cost of the compressor itself. Consequently, more efficient compressors can pay dividends in power savings over their lifetimes, even though their initial in-stalled cost may be higher.
In general, for flow requirements of 6000 cfm or less at typical shop-air pressures, reciprocating machines are the most efficient, followed by centrif-ugal (in relatively small sizes), rotary-screw, and rotary-vane machines.
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A I R C O M P R E S S O R S
Installation suggestions
Industrial air compressors are rugged machines that will perform under adverse conditions, but it al-ways is advisable to provide proper operating conditions to maximize re-liability at minimum operating cost. Traditionally, compressors have been located in separate rooms to isolate their noise. Such locations are almost mandatory today to meet OSHA requirements. However, it still is important that the compressor room have an adequate foundation (particularly for reciprocating ma-chines) as well as ample space so that the machine is easily accessible for inspection and maintenance. Stair-ways and catwalks can assist these procedures on larger compressors.
The compressor room ideally should be clean and dry. Auxiliary equipment, piping, and wiring should be arranged so that it does not interfere with routine inspec-tions. Instruments should be located within easy view of operators.
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W
hen components wear or be-come corroded as a result of moisture, they consume more compressed air — a sure sign of lost en-ergy efficiency. When this wear or corro-sion becomes great enough, components must be repaired or replaced — adding more costs to operating expense.The cost of replacement parts, labor, standby inventory, and downtime can have a devastating effect on the plant’s bottom line. Eliminating even one of the above by drying a system’s compressed air will offset the cost of installing and op-erating the equipment to do the job.
Types of dryers
Dryers remove moisture from the air, which lowers its dew point — the temper-ature to which air can be cooled before water begins to condense from it. In broadest terms, there are four basic types of industrial compressed air dryers: deli-quescent, regenerative desiccant, refriger-ation, and membrane.
Deliquescent dryers contain a
chemi-cal desiccant which absorbs moisture con-tained in the air, whether the moisture has already condensed or is still a vapor, Fig-ure 1. The desiccant is consumed in the water-removal process and must be re-plenished periodically. Solution drained from these dryers contains both liquid wa-ter and the deliquescent chemical, so dis-posal may be a problem.
Deliquescent dryers reduce the dew point of the air 15° to 25°F below the in-let air temperature. If the incoming air has a dew point of 90°F, it will leave a deliquescent dryer with dew point of about 65°F. Depending on operating conditions, some deliquescent dryers can produce dew points as low as -40°F; new deliquescent chemical may produce even lower dew points.
Two important points: desiccant level should not be allowed to fall below that recommended by the dryer manufacturer and inlet temperature should be limited to 100°F or less to prevent excessive desic-cant consumption.
Regenerative desiccant dryers
re-move water from air by adsorbing it on
duce pressure dew points as low as -50°F. The type of desiccant used has a definite effect on the final dew point.
Refrigeration dryers condense
moisture from compressed air by cool-ing the air in heat exchangers chilled by refrigerants. These dryers can produce dew points of 35° to 50°F at system op-erating pressure.
Many refrigeration dryers reheat the cooled air after it has been dried, usually by routing it through heat exchangers in contact with the hot incoming air. Reheat-ing the cooled air prevents condensation from forming on the exterior of air lines downstream from the dryer and also pre-cools incoming air.
Standard refrigerating dryers should not be used where ambient temperature can drop below 40°F because lower tem-perature can freeze condensate, which blocks air passages and could damage the the surface of a solid desiccant, usually
silica gel, activated alumina, or molecular sieve. The desiccant does not react chem-ically with the water, so it need not be re-plenished. However, it must be dried or regenerated periodically.
Heatless regenerative dryers use two
identical chambers filled with desiccant. As wet air moves up through one cham-ber, a portion of the dry discharged air is diverted through the second chamber, re-activating its desiccant. The moisture-laden purge air is vented to atmosphere. Some time later, air flow through the chambers is reversed.
Heat regenerative dryers also use
two identical chambers: one is shown in Figure 2. In this type, however, air flows through one chamber until its desiccant has adsorbed all the moisture it can hold. Then air flow is diverted to the second chamber. Heated outside air or an external source of heat (steam or electricity) then dries the desiccant in the first chamber.
Because desiccant’s adsorption capac-ity decreases as temperature increases, the desiccant bed must be cooled from the temperature it reaches during regenera-tion. The regeneration cycle in these dry-ers usually lasts several hours — 75% heating and 25% cooling.
Regenerative desiccant dryers can pro-Wet air in Dry air out Coalescing screen Collection chamber Drain solution Desiccant
Fig. 1. Deliquescent chemical dryer takes moisture-laden air into collection cham-ber, passes it though support screens into desiccant chamber where part of the water vapor is removed. ,,, ,,, ,,, ,,, ,,, Heater Desiccant Desiccant drain
Purge gas inlet Perforated stainless steel support Tubular bed support and gas diffuser Thermometer Heat-conducting fins Heater in tube Thermoswitch Pressure gauge Relief valve Desiccant fill extends into vessel to prevent overfilling
Fig. 2. Cutaway view of one tower of heated regenerative desiccant dryer shows electrically heated fins used to dry satu-rated chemical desiccant.
Liquid water in compressed air
systems can escalate a plant’s
operating expenses. Water
can damage production
ma-chinery, resulting in downtime
and spoiled product.
dryer’s evaporator. Dryers may be equipped with heat tracing packages for operating in ambient temperatures as low as 50°F.
Refrigeration dryer types
Refrigeration dryers can be further classified into three types:
Tube-in-tube refrigeration dryers,
Figure 3, operate by cooling a mass of aluminum granules or bronze ribbon that in turn cools the compressed air. As the tube-to-tube refrigeration dryer cy-cles, a thermometer in the granule mass senses its temperature. As the tempera-ture rises, a switch turns on the refriger-ation unit. When the temperature drops to a cut-off point, refrigeration stops. These dryers are designed to produce dew points of 35° or 50°F.
Water-chiller refrigeration dryers,
Figure 4, use a mass of water for cooling. An extra heat exchanger is necessary to maintain chilled water flow through the condenser, as is a water pump. Dew points can be between 40° and 50°F. Wa-ter-chiller dryers cycle as they operate.
Direct-expansion refrigeration
dry-ers, Figure 5, use a refrigerant-to-air cool-ing process to produce dew points that are 35°F below standard operating condi-tions (100°F temperature at compressor inlet, 100 psig, 100°F ambient — from NFPA standard). No recovery period is necessary, so direct-expansion refrigera-tion dryers run continuously. The cost difference between cycling and continu-ous operation is difficult to calculate. Dif-ferences in power consumption for
frac-tional horsepower dryers are negligible. The cost of starting an electric motor larger than 1 hp at the onset of every cycle becomes greater than letting it run contin-uously — but unloaded — through much of the cycle. Typically a refrigerant com-pressor will cycle once or twice every two or three minutes. Thus, in-rush loads are likely to be higher than if the motor runs continuously.
Membrane -type dryers
Membrane-type dryers are gas-separa-tion devices. They consist of permeable membrane surfaces that have been spe-cially tuned to block nitrogen and oxygen molecules (air), but allow water vapor molecules to pass right through. They work as your lungs do, venting water va-por each time you exhale.
Typically this membrane material is formed into bundles of thousands of indi-vidual fibers from one end of the dryer to the other. Water vapor escapes through the walls of the fiber to a sweep chamber from where it is continually vented to at-mosphere as a gas. A fraction of the dried air is routed through the sweep chamber to continuously purge and exhaust mois-ture vapor.
Industrial-grade membranes can be used for years to dry air continuously. They respond spontaneously to any change in inlet conditions. They perform at temperature between -40° and 150°F (ambient or inlet), and handle pressures from about 60 to 300 psig. They will de-liver a consistent outlet dew point reduc-tion anywhere between these extremes.
The inlet flow rate and pressure determine the outlet dew point depression. In other words, membrane air dryers deliver a con-sistent layer of drying protection that fol-lows the rise or fall of the inlet dew point temperature, and can easily be sized to follow the ISA recommended 20°F pres-sure dew point suppression below ambi-ent. Outlet pressure dew points can also be selected as low as -50°F. Currently, flow capacities to a relatively low 50 scfm are available, but modules can be installed in parallel for higher flows.
Prefilters mounted immediately up-stream from the dryer keep out liquids and solids to allow an almost unlimited ser-vice life. Because water vapor passes right through the membrane material, it does not accumulate there, so membranes do not become saturated and never need to be refrigerated. Membranes have no moving parts to wear out. They are non-electric and suitable for most hazardous locations.
A I R D R Y E R S
Drain Dry air out Wet air in Precooler reheater Heat exchanger Refrigeration unit Liquid refrigerant Water-to-air exchanger Wet air in
Dry air out
Water chiller Refrigerant compressor Pump Separator & drain Condenser
Fig. 3. Tube-in-tube refrigeration dryer uses refrigerant evaporator to cool wet, hot incoming air. Air-to-air precooler, top, allows heat from incoming air to warm cool, dry outgoing air. This precool-ing/post-warming process boosts overall dryer efficiency. Separa-tor collects moisture condensed from the air, drain discharges it.
Fig. 4. Water-chilled refrigeration dryers use three heat exchang-ers. Precooler/heater performs same function as in tube-in-tube dryer; second water-to-air heat exchanger pumps chilled water through the exchanger counter to air flow; third heat exchanger uses refrigerant to chill water recirculating from second to third.
Separator & drain Compressor Condenser Wet air in Refrigerant-to-air exchanger
Dry air out
Fig. 5. Direct-expansion refrigeration dryer uses two heat exchangers: air-to-air precooler/heater and refrigerant-to-air evaporator.
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They require no RF shielding or protec-tion. They use no refrigerant gas or poten-tially dusty desiccants. They vent gas, not condensate so they can’t freeze up and don’t require heat tracing at very low am-bient temperatures.
They make no noise. They can be mounted in any orientation. Their low-mass components are inherently vibra-tion-resistant. Because they are static, in-ert devices, they never need service or adjustment and don’t require monitoring devices. Made of plastic and aluminum, they do not rust or corrode and don’t need painting. They have almost no pressur-ized volume, so most pressure code re-strictions do not apply.
Note: membrane gas separators will re-move other gasses too. Membrane-type compressed air dryers can reduce outlet oxygen concentrations and must not be used for breathing air.
Importance of dew point
As pointed out earlier, wet air adds to plant operating expenses through:
● repair parts ● repair labor
● product damage, and ● production downtime.
The economic advantages of reducing or eliminating these detriments of mois-ture build a strong case for installing a dryer. Once the decision to install a dryer has been reached two questions arise: how dry must the air be and what type of dryer should be used.
The most important criterion in choos-ing an air dryer is the pressure dew point that it must produce. The required dew point of an air system determines how dry the air must be and to a great extent, which type of dryer to use. Dew point varies with pressure. For example: the dew point conversion chart, Figure 6, shows that air at atmospheric pressure with a dew point of -12°F has a pressure dew point of 35°F at 100 psig. Dryer man-ufacturers may specify the dew point that a particular model can attain at atmo-spheric pressure or at a typical system pressure, such as 100 psig. If performance is specified at atmospheric pressure, use a chart like Figure 6 to find what the mini-mum dew point will be at the system’s op-erating pressure.
The required dew point varies with each application. If preventing condensa-tion in compressed air lines is the main
concern, then the lowest ambient temper-ature to which air lines will be exposed will be the controlling factor. However, for some applications, dew point require-ments will be more severe, possibly as low as -4°F at line pressure. An example might be the air used for spraying a pow-dery substance. Even the slightest trace of moisture in such air could condense and cause particles to stick together.
If all the compressed air will be used in-side a building where temperature is maintained at a stable level, then the re-quired dew point can be fixed within a few degrees. But if some or all of the com-pressed air is subjected to outdoor temper-ature variations, the required dew point can change from day to day, or even hour to hour.
Do not be too aggressive and estimate and unjustifiable margin for error. Stating a dew point much lower than that actually required wastes money. A rule-of-thumb margin for error is about 20°F maximum. Extremely low dew points may be re-quired at only a few isolated locations. If this is the case, consider using individual dryers at each low-temperature point of use to attain these low dew points. A less expensive dryer to dry the air to
less-strin-gent requirements can then be installed for the rest of the air system.
Evaluating flow capacity
An air dryer not only must dry com-pressed air to the required dew point, but also must be able to handle the required air flow without causing excessive pressure drop. Flow capacity of a dryer depends on:
●operating pressure ●inlet air temperature
● ambient air or cooling water
tempera-ture, and
●required dew point.
When any of the above conditions changes, flow capacity of the dryer also changes. Dryer manufacturers can supply performance curves that show the relation-ship of their dryer’s flow capacity to these four factors. Table 1 compares dew point, flow capacity, inlet temperature, and ambi-ent temperature for several types of dryers. Evaluating characteristics of the different types of dryers will help indicate which is best for a particular application. This is where cost finally can be considered. Pur-chase price of the dryer is only one factor to evaluate when choosing an air dryer. A del-iquescent chemical dryer, for example, has a relatively low initial cost, but its chemical
A I R D R Y E R S
-90 -80 -70 -60 -50 -40 -30 -20 -10 0 10 20 30 40 0.001 0.003 0.006 0.014 0.021 0.044 0.088 0.166 0.285 0.481 0.776 1.23 1.93 2.84 2 6 11 26 40 83 167 316 543 916 1478 2344 3678 5412 100 90 80 70 60 50 40 30 20 10 0 -10 -20 -30 -40 -50 -60 -70 -80 -90 -100Given dew point at atmospheric pressure
Moisture content, grains/ft
Moisture content, ppm (by weight)
Pressure, psi
3
Dew point at elevated pressure
3000 psig 2000 psig 1000 psig 500 psig 400 psig 200 psig 100 psig 60 psig 400 psig 20 psig 5 psig Atmospheric
Fig. 6. Dew point conversion chart assists in determination of dew point of air at variety of pressures. Moisture-content scales chart quantity of moisture contained in atmospheric air at indicated dew points.
must be replaced periodically, adding to the operating cost. This cost is offset some-what because the deliquescent chemical dryer requires no external power source.
Other dryer types may cost more ini-tially, but have lower operating costs be-cause they can run for long periods with little or no maintenance require. It should be clear then, that cost analysis should be conducted based on manufacturers spec-ifications as they relate to an individual application’s physical and economical requirements.
Installation and maintenance
Location can affect how well an air dryer performs. The site for an air-cooled dryer should be well ventilated so heat can be carried away and readily accessible to aid maintenance. The maximum ambient temperature for a refrigeration dryer is about 100° to 100°F. Higher temperature prevent the dryer from exchanging heat with its surroundings and keep it from op-erating properly. Dryers with water-cooled condensers can tolerate higher ambient temperature because they transfer heat to the cooling water instead of to the sur-rounding environment. Refrigerant dryers, whether air- or water-cooled, should not be exposed to ambient temperature below 0°F unless optional low-ambient-temperature controls are installed.
If a deliquescent dryer is used in a cen-tral compressed air system, bypass piping should be installed around the dryer to maintain air supply whenever the dryer is taken off line to add desiccant. There should also be no set of operating condi-tions that permit system pressure to drop low enough to allow high, turbulent air flow through the dryer that might carry chemicals into system air lines.
Refrigeration and deliquescent dryers should be drained regularly, depending on the volume of liquid accumulated. Most re-frigeration dryers have automatic drains, as least as an option.
Most, if not all, refrigeration dryers re-quire a prefilter to remove oil and dirt from incoming air that would otherwise coat the inside of the dryer’s heat exchanger and lower its heat-transfer capability. Regener-ative desiccant dryers require equipment that removes compressor-oil carryover that can coat the desiccant and render it useless. These dryers also require afterfil-ters to prevent downstream migration of desiccant particles.
A I R D R Y E R S
Dewpoint needs and applications
Inlet temperature and air flow demand
Pressure
dewpoint Application Heated
Deliquescent desiccant Combination units Heatless Direct expan-sion Water chiller Urea Hygro-scopic salts Water chilled desiccant (Urea) Refrig-erated desiccant (Urea) Tube -to-tube 607F 507F 457F 357F 207F 07F and lower
General plant air for tools and fixtures when ambient temp remains at 707F or higher
Same as above, but ambient temp as low as 607F Same as above, plus general instrumentation & controls, air conveying, air gages, food & chemical processing and ambient temp as low as 557F Same as above, but ambient temp as low as 457F Same as above, but ambient temp (air lines) to –207F to prevent freeze-ups Some aerospace applications and other exotic requirements for exceptionally dry air
0-1000 SCFM, inlet 907F maximum, intermittent load, 40-80 hours per week Same as above but, but 70% capacity or more constant air flow
0-1000 SCFM, inlet 917F or higher, intermittent load, 40-80 hours per week Same as 1000 SCFM, inlet 907F maximum, intermittent load, 40-80 hours per week
Same as above, but 70% capacity or more constant air flow
Ambient temperature to 1007F Ambient temperature to 1107F Ambient temperature to 1307F Inlet temperature to 907F Inlet temperature to 1007F Inlet temperature to 1107F Inlet temperature to 1507F Inlet temperature to 2007F U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U U 1
With 607F cooling water 2
With 507F cooling water
3cooling water to 70
7F4Temperatures to 120
7F
5Water cooled only
to 1507F (1207F) (1207F) 4 5 5 5 3.4 1 2 1107F to 1307F Ambient temperatures–limitations Inlet temperature–limitations
Table 1. Guide for selecting compressed-air dryers outlines major application factors which will affect specification of an approximate unit. (Note that relatively new membrane dryers do not appear on this chart.) Check marks indicate where dryers will perform.
Refrigeration
Dual tower regenerative
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I
t is estimated that the air we breathe contains about 1/2-million particles
of dust per cubic inch. This means that there are billions of contaminating particles in the four-million cubic inches of air the average 10-hp com-pressor ingests per hour. At high con-centration and high speed, the larger of these particles can be extremely harm-ful to any compressed air system by blocking orifices, eroding components, and clogging clearances between mov-ing parts.
In addition, when ambient air is drawn into a compressor, its relative humidity can, depending on weather conditions, reach 100%. As air is com-pressed and cooled, some water vapor condenses out as free water, and even with an aftercooler, part of this liquid is swept downstream into the air system. This frequently results in rusted pneu-matic tools and components, destroyed lubricants, and frozen air lines during
downward whirling pattern. Centrifu-gal force hurls the larger solid and liq-uid water particles outward, where they collect on the inner surface of the filter bowl. The particles spiral down past a baffle into a quiet chamber. The baffle prevents turbulent air in the upper bowl from re-entraining liquid contaminants and carrying them downstream.
The dry, cleaner air then follows a convoluted path through the filter ele-ment, where finer solid particles are fil-tered out. Finally, filfil-tered air passes up the center of the element and out the discharge port.
Thus, air line filters remove impuri-ties in two operations; dynamically by centrifugal force, throwing out heavier particles and entrained water; and stati-cally through the filter element, strain-ing out smaller particles.
High-efficiency coalescing filters op-erate on a somewhat different principle from air line filters. The key difference is in the element, where a fiber network is narrowly spaced to trap smaller con-taminants. The special fibers hold any liquid particle that contacts them.
Pre-filtered air enters the cylindrical element at the center, Figure 2. As air flows through the element, particles are captured by three different mecha-nisms: direct interception as particles impinge on the fibers, inertial im-paction as particles are thrown against fibers by the turbulent air stream, and diffusion as smaller particles vibrate and collide with fibers and other parti-cles. As a result, coalescing elements can capture particles smaller than the nominal size of the flow passages through the element. Collected liquid migrates to the crossing points of the fibers where larger drops form or coa-lesce. Pressure differential through the element then forces these drops to the downstream surface of the element, where they gravitate downward to the sump. The filtered air then exits through the outlet port.
low-temperature periods.
Other types of foreign matter in air lines include:
● construction and assembly debris ● carryover oil from the compressor ● impurities generated within the
com-pressed air line — such as wear parti-cles, pipe scale, rust, and
● contaminants ingested into the air
system during maintenance or through leakage passages.
All contaminants large enough to cause air system problems should be removed by filtration. Therefore, the first step in filter selection is to deter-mine the filtration requirements of the most critical components in the system. Contamination particle size is mea-sured in micrometers — abbreviated
mm. A mm is one-millionth of a meter,
or 0.000039 in. Particle-removing filter elements are rated according to the par-ticle size they will trap. For most indus-trial applications, filter elements rated at about 40 to 60 mm are adequate. When necessary, filtration down to 20 or even 5 mm or finer can be provided. Remember, however, that finer filtra-tion increases the pressure drop through the element. As micrometer-size rating varies, so does the micrometer-size and type of filter.
Most oils entrained in a compressed air stream are in the form of tiny mist or aerosol droplets that can pass through standard industrial filter elements. A coalescing-type filter can be used to re-move these aerosols. The sub-microme-ter oil particles that escape a coalescing filter should have no detrimental effect on industrial pneumatic components. However, if these particles must be re-moved for applications such as food processing or breathing air, a third oil-removal-type element is available.
Filter operation
When pressurized air enters a typical filter bowl, Figure 1, a curved inlet and deflector direct the incoming air in a
Reliability is one of the strong
suits of pneumatic systems —
and proper filtration is a key to
reliability and longevity. Here
are some basics that explain
the whys and hows.
Air filters
Fig. 1. Cutaway drawing shows internal workings of typical airline filter. Baffles di-rect air into circular motion where centri-fugal action separates contaminants, which collect in quiet zone, from where they can be removed.
, Wet air
enters Dry air exits
Cleanable porous bronze element Polycarbonate bowl Metal guard Quiet zone eliminates carryover Multi-stage
baffle spins air to separate moisture
