Hydrocarbon Compression


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The form of a purchase specification should be the one most familiar to,
and most commonly used within, the issuing organization; typical forms
are given in API standards 617 and 618. Here we shall consider the content
of the specification.

Performance criteria must be carefully defined for the end use that the compressor must have within the overall system. Following should be included in the specification:
• The range of mass and volumetric flow as influenced by variations in the inlet temperature, pressure, molecular weight, gas composition (vapor loading, compressibility factor, etc.), discharge pressure, temperature, and flow of cooling fluids (water, air, etc.)
• Startup, standby, and shutdown conditions of the compressor and of the entire system
• Mention of even traces of vapors, liquid droplets, dusts, or gases that may be minor items for the chemistry of the process but may cause fouling, gunking, seal problems, etc., either by themselves or when mixed with lubricants or sealing fluids (items such as these may appreciably influence the choice of the compressor type)
• Range of ambient temperature
• Altitude
• Area electrical classification
• Applicable codes and standards from such organizations as the Tubular Exchanger Manufacturers Association (TEMA) and the American Society of Mechanical Engineers (ASME)

Purchase specifications must define quality requirements for auxiliary equipment such as seals, piping systems (material and arrangement), type and quality of control elements and systems, level of redundancy, and shop testing (if any). Checklists for such items may be prepared based on the available knowledge within an organization, as well as on accepted references on the inspection of completed installations, such as Chapter X of the API “Guide for Inspection of Refinery Equipment.” Purchase specifications—or those prepared by the customer for use by an engineer-constructor—should not limit the bidders from using their knowledge and experience.

Required controls cover a very wide range of supply. The compressor specification should include all elements that directly measure and control any part of the compressor system. This includes local panels, receivers from external inputs, and any items to provide outputs to external devices. Devices for volume control as such—or those used to control mass flow and provide anti-surge control—must be carefully defined as to which elements are supplied as part of the compressor system and which elements are external. Thus, such items as inlet guide vanes are best included in the compressor system, while anti-surge and recycle devices are usually best considered as external to the system.

An increasing area of interest for controls consists of the types of diagnostic devices used to measure, indicate, alarm, and record vibration (velocity and displacement), axial movement, bearing temperatures, and drive-motor-copper temperatures. Axial-movement and motor-copper temperature indicators are best used for both alarm and shutdown. Other instruments are most suitable for alarm only and as trend indicators. The compressor supplier is in the best position to select the points of pickup and recommend types of pickup and readout devices. Controls for units to be attended only by remote or occasional local surveillance require very careful attention.

Job cost and completion time is improved with proper use of shop-assembled units. Typical packages, including skid-mounted units, comprise refrigeration (chilled water and low-temperature brines) and complete instrument and plant-air units. Purchase specifications should therefore call for or permit packaged units to be offered where feasible. The units should be such that they need merely to be set on simple foundations, and have the power, cooling-water, and supply and discharge piping connected.

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October 1st, 2021 at 12:59 am

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Smaller units are usually electric-motor driven (direct or belt); for medium to large units, a wide choice of drives are available. These include motors (synchronous, induction, low or high speed); steam turbines (back-pressure, condensing or controlled extraction); internal-combustion engines (integral or direct connected); gas turbines (single or double shaft); and expanders.

The selection of a drive depends to some extent on the compressor service, but more important are the overall energy balance, energy utilization and availability, and heat-rejection methods. Within the limits imposed by these criteria, the selection should stress a drive system that is simple, dependable, and straightforward. The compressor is the reason for the drive, not the other way around.

The drives of internal-combustion engines and steam turbines can ordinarily be operated over a fairly large range of speeds. This may not be the case, however, if gas-turbine or electric-motor drives are used.

Let us first consider gas turbines, almost all of which have axial-type air compressors for pressure–air supply to the compressors. For single-shaft units (air compressor, gas turbine, and driven unit on one shaft), the speed range is most often determined by the steep performance curve of the axial compressor rather than by the much flatter curve of the centrifugal process compressor. Double-shaft machines permit constant speed for the axial air compressor and variable speed for the process compressor. The selection of sizes, speeds, and horsepower outputs of commercially available gas turbines is limited. Very often, a wide freedom of choice is not available as to single- or double-shaft units.

Motor drives usually have a constant speed: In a limited number of cases, variable-speed couplings, wound-rotor or multipole (PAM) motors may be used. Large motors may be of the synchronous or induction kind. For a unit driven at above-synchronous speeds (3,600 rpm for 60 Hz), the choice should be based on the total cost of the motor and the speed increaser. Thus, a 1,800-rpm induction motor and its speed increaser may cost less (including operating costs) than a 1,200-rpm synchronous motor with its speed increaser.

Constant-speed centrifugals in process plants tend to operate at a high enough average load so that the economic rewards of power-factor correction obtained by the use of a synchronous motor are minor.

Fossil-fuel drives are used when initial and operating costs are more attractive than those of steam or motor drives, sufficient electric power is not available, and electric or steam sources are not reliable. In this last case, the entire system must be carefully specified to ensure that minor items such as cooling-water pumps, pressure switches, control air, etc. are independent of any source of power, not as dependable as the compressor drive.

Internal-combustion engines are usually turbo-supercharged and may be two or four cycle, integral with or separate from the compressor. The type of engine can normally be selected on the basis of drive features, including accessories and costs (purchase order, installation, fuel consumption, spare parts, maintenance) independent of the compressor. It is best, however, to include the drive as part of the compressor system.

The gear mechanical rating, including the American Gear Manufacturers Association (AGMA) service factor, should be selected so that the gear rating does not become the limiting factor in the compressor and drive train. Steam-turbine drives combined with a gear (with the turbine at a lower speed than the compressor) are sometimes lower in cost than higher-speed turbines. The policy of the user and his insurance carrier on warehouse spares for gears affects the choice, since the gears and additional couplings increase the probability of outage.

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October 1st, 2021 at 12:49 am

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At times, combining compressors may be worth considering. For example, in compressing to a very high pressure, it may be possible to use a centrifugal machine or a rotary-screw machine for a lower pressure and then pipe the gas to a reciprocating unit. In some instances, axial and centrifugal impellers may be placed on the same shaft. In addition, one might also resort to placing axial- and centrifugal-compressor cases in a common drive train.

As an alternative to an axial compressor, three or four single-stage centrifugal compressors may be connected by a gear train to a single drive. With the gas cooled after each stage of compression—and gears designed to permit each stage to be run at its optimum speed—the efficiency of these centrifugals is comparable to that of an axial compressor, while their operating characteristics are those of a centrifugal machine. Extensive gearing, however, is a distinct disadvantage.

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October 1st, 2021 at 12:41 am

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The type of mechanical drive (including gears) that is used may influence the choice of compressor.

Compressor and drive speeds are very pertinent if one wishes to avoid gearing. The accompanying table 1.1 provides speed ranges of the most common types of compressors and drives. There are specially designed units, however, that do not fall within the ranges listed. One of these, for example, is a carbon dioxide compressor with a suction volume of approximately 50 acfm at the last wheel, which rotates at 25,000 rpm and delivers gas at 5,000 psi. The tip speed of this compressor’s impeller is approximately 650 ft/sec. The compressor itself is directly driven by a specially designed 1,000-hp steam turbine.


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October 1st, 2021 at 12:30 am

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A positive-displacement compressor is characterized by a pressure rise volume curve that is almost vertical. (It is not completely vertical because there is mechanical clearance, and slip and leakage from the discharge to the suction; the slip increases as the compression ratio rises.) The compressor delivers its gas against any pressure head up to the limit of its mechanical strength and drive capacity. Capacity is almost directly proportional to speed.

The characteristics of a centrifugal compressor are appreciably different. Generally, the pressure rise–volume curve is quite flat (Figure 1.4a). (It may be somewhat steeper if a heavier gas is being compressed.) A small change in the compression ratio produces a marked effect on the compressor output. As the discharge pressure increases, the flow ­decreases, and if the flow decreases too much, the machine will start to surge.

Surging occurs when the velocity of gas leaving an impeller wheel is too low to move the fluid through the machine. With no gas leaving the impeller, the discharge pressure may drop. Should this occur, the machine will again start to compress gas, and the cycle will be ­repeated. Such intermittent operation may severely damage a compressor. The characteristic curve can be modified by the installation of adjustable inlet guide vanes (Figure 1.4b). These are most effective on machines having a few stages. Adjustable diffuser vanes have been used on some machines.

In some installations, process requirements may dictate that the compressor be run at the far right of the characteristic curve, where it is very steep. Operating in this area requires careful control and is accomplished at some penalty of compressor efficiency.

The volumetric capacity of a centrifugal compressor is almost directly related to its speed; its developed head, to the square of the speed. (The horsepower requirements are thus related to the cube of the speed.) The efficiency of centrifugal compressors is lower than that of reciprocating machines by perhaps 5 to 20 percent.

These characteristics establish the sensitivity of the compressor to variations in flow conditions. For example, a change in the density of the fluid being compressed will have little effect on either the volume of gas pumped or the discharge pressure developed by a reciprocating machine, although one would have to be sure that no component parts of the compressor were being mechanically overstressed. Any variation in the density of a gas being compressed will result in a proportionate change in the weight of gas pumped.

On the other hand, because the head developed by a centrifugal compressor depends only on the velocity developed, a change in gas density will be directly reflected by a proportionate change in the developed discharge pressure. However, at a given density, if the discharge pressure can be permitted to change slightly, one can obtain large variations in volumetric gas flow through the compressor.

The axial compressor has a very steep characteristic curve (Figure 1.4c). The unit’s surge capacity is thus close to its operating capacity. However, by providing a method of adjusting the angle of stator blades and inlet guide vanes, a greater operating range can be obtained (Figure 1.4d).

Generally, the efficiency of an axial compressor exceeds that of a multistage centrifugal machine by perhaps 5 to 10 percent. The axial compressor does not contain diaphragms that expand radially as the compressed gas gets hot. This mechanical factor, combined with higher efficiency, leads to greater freedom from temperature limits and permits a higher compression ratio per case than do centrifugal units.

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October 1st, 2021 at 12:25 am

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Not all types of machines are made in all pressure–volume ranges. Figure 1.1 indicates, in a very general manner, the capacities of reciprocating, centrifugal, rotary-screw, and axial compressors available. The more common usage is indicated by the deeper shading. Although this figure does not indicate the theoretical or engineering limits of any design (the limits are continually being expanded), it may be used as a guide to current technology.

Since sealing systems for axial compressors are not as versatile as for other types, normally only those gases whose leakage to the atmosphere can be tolerated should be handled by this type of machine.

Rotary-lobe, sliding-vane, liquid-piston, and diaphragm compressors have relatively small capacities and, typically, atmospheric-pressure suction. Of these four types, the rotary-lobe compressor can deliver the most gas, as its maximum suction volume is about 30,000 actual cfm (acfm). A maximum discharge pressure of about 40 psia can be attained. However, rotary-lobe compressors are most competitive at capacities of 17,500 acfm or less and discharge pressures of about 22 psia.


Maximum inlet capacities of sliding-vane units are about 3,000 acfm, or double this amount if a duplex compressor is used. The latter consists of two compressors attached to a single drive. Maximum discharge pressures of standard machines are about 65 psia in a single stage and 140 psia in two stages.

The liquid-piston compressor has a maximum capacity of about 10,000 acfm and can deliver this amount of air (or gas) at about 30 psia. Volumes of 300 acfm or less can be compressed to about 115 psia. The three foregoing compressor types can produce moderate to high vacuum, particularly in multiple stages.

Diaphragm compressors have much smaller volumetric capacities, with maximum flows ranging from 40 to perhaps 200 acfm. These machines, however, can develop pressures up to 40,000 psi. Before selecting a compressor type, one must decide how many machines will be needed to handle the process load. In former years, reciprocating machines were used for almost all process applications. Since the compressor capacity was low, large plants would require trains of machines. As machine reliability and capacity increased, the tendency to install two machines started, each with 55 percent or 60 percent capacity, perhaps with a third unit as a spare.

The spare unit ensured operation at full capacity, but at an increased compressor cost of about 50 percent. If the spare compressor were omitted, but two half-size machines installed, one could still be reasonably sure of continued operation at all times. This was particularly important when the process included equipment that could not be shut down frequently, such as furnaces.


Later, to take advantage of larger machine capability, several services were placed on the same frame.

Today, the situation is somewhat different, as more centrifugal compressors are being used (Figure 1.3). For one thing, the downtime of a rotating equipment generally is appreciably less than that of a reciprocating equipment. Therefore, in many instances a single centrifugal compressor may be satisfactory. However, it must be recognized that when a compressor is down, it will usually take longer to repair or overhaul a centrifugal unit than a reciprocating one—unless a complete spare rotor is available.

Also, the pricing structure of centrifugal compressors is quite different from that of reciprocating ones. As a first very rough approximation, one may assume that halving the size of a reciprocating compressor will halve its cost. Yet, halving the size of a small centrifugal compressor may only decrease its cost 20 percent, and halving the size of a large machine may only reduce its cost 30 percent.


Furthermore, because of their flat operating characteristics, the running of centrifugal compressors in parallel may result in surging unless very careful attention is given to avoiding unstable operation. Therefore, in many process applications for which one centrifugal compressor will have adequate capacity, an installed spare is not provided. In these instances, a complete spare rotor may be bought.

The choice between reciprocating and centrifugal compressors is not always simple, particularly for high-head, medium-capacity service such as gasfield repressuring. If several reciprocating compressors are used, each can be multiple staged to develop the desired head. The shutting down of one machine would merely cause a decrease in plant output. But if several centrifugal compressors were used in series, the failure of one would stop the entire operation.

Written by Jack

October 1st, 2021 at 12:19 am

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As noted earlier in this chapter, when you are moving toward selecting a rotary-screw (or any compressor type), you first need to know how much air you’ll need in cubic feet per minute (CFM) at the psi you need for your plant, your tools, and all ancillary equipment for now and for the future. You’ll want to build in excess volume available, as one statistic we’ve seen says, on average, more than 10 percent of all compressor capacity is lost through leaks, despite the best efforts of the plant to reduce wasted air.

Once you got the compressor size figured out (link; I hope the information here will help), you will want to review the up-front cost of compressors from a host of manufacturers. Check their mean time, between-failure rates, their parts and service costs, the life expectancy of the unit with the duty cycle you will require, and the particular operating costs.

Compressing air is expensive, and one compressor might provide lower up-front capital costs, yet end up being far more expensive in the longer term due to higher operating costs.

All factors having been considered, and certainly this is claimed by many of the manufacturers of the rotary-screw type of compressor, the rotary-screw compressor may surface as your best choice for industrial application.

• Sliding-vane compressors, where an eccentric cam (into which sealing vanes slide) rotates inside a housing.
• Liquid-piston type, in which a partially liquid-flooded case creates the equivalent of sliding vanes.
• Diaphragm compressors, in which a flexible diaphragm is pulsed inside a concave housing.

The two types of compressors that convert velocity into pressure are:
• Radial-flow compressors, generally called “centrifugal compressors”
• Axial-flow compressors, known as “axial compressors”

In centrifugal compressors, the gas enters the eye of the impeller, and the rotative force moves the fluid to the rim of each wheel or stage. Diffusers convert the velocity head into pressure, and return passages are then used to lead the gas to the compressor discharge or to the next impeller stage.

In axial compressors, flow occurs through a series of alternating rotating and stationary blades, and in a direction basically parallel to the compressor shaft. Each passage through the rotating blades increases the velocity of the fluid, and each passage through the stationary diffuser blades converts the velocity head into a pressure head.

Written by Jack

October 1st, 2021 at 12:12 am

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Although rotary-screw compressors are available in lower horsepower ratings, it would appear that it is in the 20 to 25 horsepower and higher range that many industrial compressor applications tend to move toward using a rotary-screw compressor solution instead of other types of compressors.

One major manufacturer states that the rotary-screw air compressor has become the most popular source of compressed air for industrial applications.

That may be because of the need for a compressor with a continuous duty cycle. Some rotary-screw compressor manufacturers claim a duty cycle of 24/7/365, which is pretty incredible for any electromechanical device.

Rotary-screw compressors are available with a direct motor-to-screw drive; others are belt driven. Each has its benefits and its own drawbacks, the details of which are best obtained from the actual compressor manufacturer.

Less Maintenance
The perception, one that is claimed by some manufacturers, is that rotary-screw units have the least maintenance issues of all types of air compressors and are touted as being the easiest to maintain in terms of both regular maintenance and unscheduled downtime.

Reputed for Lower Cost
When you move up into the higher horsepower units, rotary-screw units are reputed for their lower cost over a comparably sized reciprocating compressor, and further, they boast lower operating costs than either vane or reciprocating units.

Oil Carryover
Some manufacturers suggest their oil carryover from the compressor to the compressed air supply of the plant is calculated in parts per million per day, rather than the ounces or more of oil that can enter the plant-air stream from older reciprocating models and some well-used vane models.

Lower Operating Noise
Other firms suggest that their rotary-screw units boast a very low operating decibel rating, and claim noise output levels far below other types of compressors, an important issue to be considered for the benefit of workers in the area. It is our experience that the lower operating sound levels may not eliminate the need for a soundproof housing, unless the compressor is well equipped with sound-deadening cladding.

Written by Jack

October 1st, 2021 at 12:11 am

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If you have ever used an electric drill in a piece of wood or metal, you may have noticed that the chips or spirals of metal follow the contour of the flutes up and out of the hole (most of them, at least).

A similar phenomenon occurs inside the rotary-screw compressor housing.

At the wide end of the screw (sometimes there is one screw operating against a housing, sometimes more than one), an inlet valve allows free air into the screw housing when there is a demand. Free air flows into the housing from the outside as there is a partial vacuum formed inside the rotary-screw housing as the screw(s) rotate.

Inside the screw housing are the screws in a bath of oil. The oil is there to provide a viscous, flowing, sealing method to help trap the air in the rotary-screw flutes.

The air–oil mixture in the screw housing moves along the flutes from the wide end of the screw toward the narrow end, pulling a vacuum behind, thus drawing more air into the screw housing.

As the air–oil blend is pulled along the flutes of the screw, the space in which the air is contained gets smaller and smaller. The diameter of the screw is larger at the inlet end and smaller at the discharge end, thus compressing the air. The amount of air trapped in the screw flutes does not change as the air is moved along the narrowing path, but the volume that air is in gets steadily smaller, thus compressing the air.

Manufacturers of rotary-screw compressors have their own ideas of what constitutes the ideal geometry of the screw within their air compressor.

Rotary-screw compressors may have just one screw (also sometimes known as augers) or maybe two or more. Single-screw compressors function the same way as multiple-screw units, with the air being compressed between the housing of the screw compartment and the screw ­itself, rather than between two or more screws.


The following drawing will give you an idea of how the rotary-screw concept works with two screws. The actual guts of the rotary-screw compressor will vary depending on the designs of the company that manufactured that particular compressor. The drawing shows two screws. They would be housed inside the screw compartment of the compressor, in a bath of oil.

At the narrow end there would be an outlet valve, which feeds the compressed air–oil mixture from the screw compartment and into a separator.

The separator has the job of removing as much oil from the compressed air as possible, and then to release that compressed air into the compressor receiver or into the plant main air lines.

Written by Jack

October 1st, 2021 at 12:08 am

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Rotary-screw compressors can fail in a number of ways. The most common problem is oil in the compressed air. Most of the times this happens because the oil separator is not doing its job properly. The chances are that the separator element is saturated with oil (read: you didn’t service the compressor in time! There’s a fixed time limit [running hours] to change the element!).

Another problem often encountered is water in the compressed air. Since the compressor takes in a huge amount of air (with water vapor) and compresses it to seven times a smaller volume, a lot of water will be produced. Normally this water is drained using an electronic or mechanical automatic drain. If this drain is broken, the water will stay in your compressed air and fill up your air receiver and piping.

If the problem persists, it will most probably be a defective pressure switch (which will start/stop, load/unload the compressor), or a defective inlet valve (which opens and closes the air inlet of the compressor). If it’s closed, the compressor is running in an “unloaded” condition and won’t supply any air.

Written by Jack

October 1st, 2021 at 12:05 am

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