Centrifugal & Axialcompressors

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Unit P-10–03

Compressors

Centrifugal &

Axial

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Centrifugal & Axial Compressors - Unit P-10-03

UNIT P-10-03

CENTRIFUGAL AND AXIAL COMPRESSORS

TABLE OF CONTENTS

1.0. OBJECTIVES/INTRODUCTION..........................................................................4

2.0 CENTRIFUGAL COMPRESSOR THEORY............................................6

2.1 Centrifugal force:.......................................................................................6

2.2 Centrifugal Force and Diffusion...............................................................7

2.3 Compressor Casing....................................................................................8

3.0................................................................ Horizontally Split Casing

9

4.0................................................................... Vertically Split Casing

11

4.1 Impellers....................................................................................................12

4.2 Diffuser, Volute.........................................................................................13

4.3 Shaft and Rotor........................................................................................14

4.4 Balance Drum............................................................................................16

4.5 Bearings.....................................................................................................17

4.6 Shaft Seals................................................................................................18

4.6.1 Labyrinth Seal...........................................................................................19

4.6.2 Restrictive carbon rings............................................................................20

4.6.3 Mechanical Face Seal...............................................................................21

4.6.4 Liquid Film Seal........................................................................................22

4.6.5 Dry Gas Seals...........................................................................................23

4.7 Diaphragms...............................................................................................25

4.8 Guide Vanes..............................................................................................26

4.9 Couplings...................................................................................................27

5.0.......................................................... COMPRESSOR LUBRICATION

28

5.1................................................ Lube Oil System Flow Description .

31

5.2 Lube Oil System Instrumentation..........................................................32

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6.0................................................. COMPRESSOR OPERATING CURVES

33

6.1 Compressor Control.................................................................................37

6.2 Capacity Control.......................................................................................40

6.2.1 Capacity Control of Single Compressor Units...........................................47

6.3 Surge Control............................................................................................50

6.4 Stonewall...................................................................................................53

6.5 Gas Density...............................................................................................53

6.6 Critical Speed............................................................................................54

7.0 COMPRESSOR MONITOR AND PROTECTION DEVICES.....................55

7.1 Alarm and Shutdown Protection............................................................55

7.1.1 Start-up/Shutdown....................................................................................55

7.1.2 Monitoring System....................................................................................56

7.1.3 Mechanical Failure Alarm and Shutdown..................................................58

8.0 AXIAL COMPRESSOR DESCRIPTION..............................................59

8.1 Construction Materials..........................................................................61

8.2 OPERATING PRINCIPLES..........................................................................62

8.3 Associated Problems...............................................................................64

9.0 LUBE AND SEAL OIL SYSTEM.......................................................65

9.1 Lubricant......................................................................................................65


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1.0. OBJECTIVES/INTRODUCTION

Objectives

At the end of this lesson the trainee will be able to:

Describe the operation of a centrifugal compressor.

List the types of prime movers that are used for centrifugal compressors.

Name and describe the main parts of a centrifugal compressor

Describe methods of capacity control used on a centrifugal compressors

Describe the monitoring, alarm and shutdown systems

Identify the main parts and describe the construction of an axial compressor

Describe the operating principles of an axial compressor

Describe surge and stall conditions


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Introduction

Early applications of centrifugal compressors were mainly for low pressure/high flow rate applications however they are now widely used for duties previously handled by reciprocating compressors.

Centrifugal compressors have a number of advantages compared with reciprocating machines. For a given duty the centrifugal compressor will tend to be less expensive and more compact than its reciprocating counterpart. Centrifugal compressors have the added advantages of providing a non-pulsating discharge flow and of delivering oil free gas. Due to their design centrifugal compressors are more reliable than reciprocating compressors.

Centrifugal compressors compress gas by the mechanical action of rapidly rotating impellers or bladed rotors that impart velocity and pressure to the flowing gas. Velocity is converted into pressure in stationary diffusers.

Most homes will have a simple axial compressor, in the shape of a cooling fan. This fan will comprise of a few blades mounted on to a hub that rotates. The shape of the rotating blades causes air to be sucked in at one side and blown out of the other.

Axial compressors are commonly used as part of a combustion gas turbine for compressing atmospheric air prior to combustion with fuel gas before -the hot expanded gases pass through the turbine. Axial compressors are named because of the gas flow path through the machine, which is parallel to the shaft (or axis) as opposed to the radial flow in a centrifugal compressor.

The centrifugal flow compressor can be either single or multistage using an impeller to accelerate the gas and a diffuser to produce the required pressure increase. The axial flow compressor is always a multistage unit employing alternate rows of rotating (rotor) blades and stationary (stator) blades, to accelerate and diffuse the gas until the required pressure rise is achieved.

Although the centrifugal compressor is easier to construct and is more robust, the axial compressor can handle much more gas for the same size of compressor. It is also more efficient because the gas flow is kept at subsonic speeds resulting in an almost straight through path, where as gas flow through a centrifugal compressor occasionally exceeds supersonic speeds in places causing shock waves.

Also, by the nature of its design, the gas flow in a centrifugal compressor has to change direction at right angles a number of times creating a high amount of friction. Axial compressors have a smaller frontal area, which is an important consideration for aero engines from which the industrial turbine is adapted. However, one drawback is that the operating range is quite small.

An axial compressor consists of a number of rotor assemblies that carry blades of an aerofoil section i.e. they have a curved shape similar to an aeroplane wings’ cross-section. The whole rotor is supported by bearings in a casing, which contains the stator blades.


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2.0 CENTRIFUGAL COMPRESSOR THEORY

Centrifugal compressors are a type of dynamic compressor that uses two basic principles; centrifugal force and diffusion, to add pressure to a gas.

2.1 Centrifugal force:

This type of force can be demonstrated by pouring a liquid into a container. If the container is then swung around in a circular motion, the liquid is held in the bottom of the container by centrifugal force. If a pressure gauge is fitted to the bottom of the container, it shows a pressure greater than the liquid head only.

If the liquid is replaced by a fluid with a lower density such as gas, it would react in the same way but the pressure shown on the gauge would be much less due to the lower weight of gas.

Diffusion (velocity to pressure principle):

Energy cannot be created or destroyed but it can be converted from one form to another. This can be demonstrated by the actions that occur when a gun is fired. The chemical energy of the burning gunpowder is converted and transferred to the bullet giving it kinetic energy and the bullet leaves the gun barrel at high speed. When it strikes an object, the kinetic energy is converted into pressure energy.

The two principles (centrifugal force and diffusion) combine to give the total increase in pressure in a centrifugal compressor.

Of these two principles, centrifugal force provides approximately one third of the total pressure, while the velocity to pressure conversion accounts for two thirds. This is shown in Figure 2-2.


Figure 2-1

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2.2 Centrifugal Force and Diffusion

Centrifugal compressors take mechanical energy from a driver e.g. an electric motor via a drive shaft. Connected to the drive shaft is a device called an impeller. This rotating impeller accelerates the fluid by subjecting it to centrifugal force.

The fluid, now at high speed, leaves the impeller’s outer rim and passes into a part of the compressor casing called a diffuser or volute. The diffuser/volute reduces the gas velocity of the gas which results in an increase in gas pressure.

The impeller is equipped with several blades or vanes to maximise conversion of the mechanical energy from the driver into kinetic energy. The gas is contained and directed in an enclosure called a compressor casing that also contains the parts that cause a cause the gas to rise in pressure.


Figure 2-2

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2.3 Compressor Casing

A simple casing, containing a single impeller, is generally found in low-pressure machines. This type of casing is easy to access for repair and maintenance. Usually, unbolting a centre cover reveals the impeller and a single bearing.

A single centrifugal impeller is limited in the amount of pressure increase obtainable. For example, if a gas has an initial inlet pressure of 0·15 barg, a single impeller compressor is capable of increasing that pressure to about 0·35 barg.

If a higher discharge pressure is required, several impellers can be placed in series within the casing. Each impeller stage takes its suction from the discharge of the previous impeller. This arrangement of a series of impellers is called a multi-stage compressor.


Figure 2-3

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3.0 HORIZONTALLY SPLIT CASING

For multistage casings designed for low to medium pressure ranges (1·5 - 50 barg) most manufacturers use a horizontally split casing arrangement.

With this type of casing, the top half can be easily removed. The casing is built to split apart at the horizontal centre line, the top and bottom halves being held together by bolts and positioned by dowel pins. By lifting the upper casing, all compressor internals, such as the rotor, labyrinths, seals, bearings, and diaphragms, can be accessed for inspection and repair.

Figure 3-1

Figure 3-1


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These casings can be manufactured in several ways with the intake, discharge, and side-load nozzles positioned on top, sideways or bottom of the casing. Side-load nozzles between the inlet and discharge allow intake or removal of gas at different pressures. Several casings may be driven by one driver; the shaft being extended through both ends of the casing where necessary. This arrangement is commonly called a Tandem System.

Stress, temperatures and corrosion determine materials used for horizontally split casings. For normal temperature ranges of -10 to 130C, cast iron, cast ductile iron, or cast steel of varying grades are used. API Standard 617 requires steel casings for the following conditions:

Gas pressure above 20 barg.

Flammable or toxic gas at a pressure over 5 barg.

Gas gaining heat of compression in excess of 150C.

Flammable or toxic gas gaining heat of compression heat above 115C.

For high stress or low temperature conditions, alloys are available. Where corrosion is expected, the casing can be made thicker than normal. When severe corrosion is expected, corrosion resistant alloys such as stainless steel and chromium are used in both the casing and internals.

Casing (belly) drains are installed at the bottom of each impeller volute. These are simple tapped, flanged or threaded openings connected to a gate valve. The horizontally split casing flanges have a pressure limitation of about 50 barg.


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4.0 VERTICALLY SPLIT CASING

A vertically split type casing (sometimes called ‘barrel’ type) is used for high-pressure services. It is a forged or cast steel cylinder made to withstand higher internal pressures.

Figure 4-1

With this design, changes in metal dimensions caused by temperature rise are symmetrical and relatively equal, eliminating large stresses and distortion encountered in other types of compressor casing. Heavy forged end covers can be fitted to both ends or just the outboard end depending on the design.


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4.1 Impellers

As mentioned earlier, the impeller is the rotating element in a compressor that increases the gas velocity. They come in many shapes and sizes. They are classed as open, partially open or closed, and by the direction of vane curvature and method of construction (cast, milled, riveted and welded).


OPEN PARTIALLY OPEN CLOSED

CURVED FORWARD

CURVEDBACKWARD STRAIGHT

Figure 4.2

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Vanes can curve forward, be straight (called radial), or curve backwards. Most compressors use backward leaning vanes because they are more efficient and more stable, but will produce a slightly lower pressure increase than other types under the same operating conditions. Oilfield compressors generally use backward curving closed impellers.

The diameter of an impeller is limited by the amount of stress generated by the high speeds under which they operate.

Impellers undergo material testing and dynamic balancing (spin test) during manufacture and after machining (assembly of rotor). The speeds at which they are tested are usually 100-115% of maximum operating rpm.

4.2 Diffuser, Volute

As the gas leaves the impeller tip it is forced into a passageway in the casing called the diffuser.

The diffuser forms a flow path that directs gas to the discharge nozzle of the compressor. Once the gas is in the diffuser its velocity decreases. This decrease in velocity converts kinetic energy of a gas into an increase in gas pressure.

Figure 4-3


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For a single stage compressor and for the final stage of a multistage compressor, the diffuser spills gas into a spiral shaped passageway called a volute. The increase in area within the volute slows down the gas velocity even more and the gas pressure is again increased.

Each stage of a multistage centrifugal compressor has all the parts of a single stage compressor. One stage is separated from the next by a specially formed part of the casing called the diaphragm.

4.3 Shaft and Rotor

A multi-stage compressor has two or more impellers on a single shaft. The shaft including sleeves, impellers, balance drum, etc. is called a rotor. All of these items are fixed to the shaft and rotate inside the compressor casing.


Figure 4-4

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Figure 4-5

Shafts are expensive to manufacture as high quality metal must be used that is capable of being machined to close tolerances. To reduce cost they can be made of mild steel and then plated with chrome, stainless steel or other hardened metal to protect the shaft.

Parts of the shaft exposed to wear or corrosion (such as the area around the seals) are often protected with a removable sleeve made of special alloys.

At high speeds, vibration of the rotor inside the casing can cause severe damage. Although the rotor is carefully balanced during manufacture, it will still vibrate at certain speeds. These vibration points called critical speeds can be predicted, the normal operating speed is set to avoid these points.

On start-up and shut down, compressors must pass through their critical speeds as quickly as possible. It is important never to hold a compressor or turbine at its critical speed for any length of time.

Figure 4-6

Flow through an impeller is unbalanced. A multi-stage rotor will have extreme forces pushing it towards the suction end. This force is called thrust.

There are several ways to minimize thrust force. A back-to-back rotor arrangement is used to reduce the net thrust acting on the thrust bearing, see Figure 4-7. With two sets of impellers arranged back-to-back, the greater part of thrust is balanced by the impellers themselves.

Figure 4-7


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Another method is to machine spiral ribs on the back of the impellers. These ribs cause a ‘pumping’ action that counteracts thrust forces.

4.4 Balance Drum

A centrifugal compressor rotor is normally built with a balance drum or piston installed at the discharge end to relieve the thrust bearing load. This is the most common method used to balance opposing forces on the shaft. A metal cylinder is fitted to the shaft just behind the last impeller.

BALANCING PISTON

IMPELLER

SHAFT

Figure 4-8

Full discharge pressure acts on an area on one side of the balance drum, while compressor suction pressure is piped to an area on the other side. Proper sizing of the balance drum allows control of the direction of thrust by balancing the resultant forces.


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Figure 4-9


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4.5 Bearings

Thrust bearings

Thrust bearings are limiting device that restrict shaft movement caused by axial thrust. With low thrust forces, a tapered roller bearing can be used. In larger machines, a heavy-duty tilting-pad type bearing is used.

Figure 4-10

A solid metal thrust collar or thrust disc is fixed to the shaft. Four to eight machined metal pads set in sockets that allow the pads to tilt in any direction. The pads are positioned in a circular bearing case. Depending on thrust conditions, one or two tilting-pad bearings can be used.

Figure 4-11

If two thrust bearings are needed, one is placed on each side of the thrust collar. The bearings are attached to compressor casing and do not rotate with the collar. The pads position the thrust collar and the shaft to limit the axial movement of the shaft in either direction.


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Journal or Main Bearings

Journal bearings support and position the shaft and allow it to turn with the minimum of friction.

Several types may be used such as self-aligning sleeve, ball or roller. In low pressure machines the bearings are often installed on the outside of the compressor casing to provide easy access for inspection or maintenance.

Large compressor units use a plain bearing. The inner surface of the bearing is bonded with a soft metal to minimise wear on the shaft surface.

In high-speed units, an operating problem known as oil whip may occur. Lube oil pumped into a bearing tends to build up a mass on the shaft away from the load, causing rough operation and increased bearing wear.

Special grooves cut into the shaft or liner create ‘dams’ that force the oil to all parts of the bearing. In some cases an anti-whip bearing is used, this type of bearing has tilted pad shoes to prevent the oil mass from forming and distributes oil evenly over the bearing surface.

Figure 4-12

4.6 Shaft Seals

A seal is a device that prevents a higher pressure fluid from leaking into a lower pressure fluid. A compressor requires a sealing device to stop the compressed gas from leaking to atmosphere (or to keep air out of vacuum compressors) and to prevent interstage leakage between impellers.

Complicated sealing systems are required when the compressed gas is combustible, toxic, corrosive or when it must not be contaminated with air. The main functions of seals are to isolate the gas inside the compressor from the atmosphere and to prevent high pressure gas from passing back through the compressor internals.


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The most common types of seals are:

Labyrinth

Restrictive (segmented) carbon rings

Mechanical face (metal contact)

Liquid-film

Dry gas seal

Seal selection will depend upon the compressor service.

4.6.1 Labyrinth Seal

The labyrinth seal is the simplest and most common type of seal used on centrifugal compressors. A series of sharp teeth are machined on the inside of a circular metal liner fitted to close tolerances around the shaft.

Machining teeth on the shaft that interlock closely with those on the liner can reduce the amount of leakage however some leakage from the high-pressure side to the low-pressure side is allowed.

The seal teeth are machined on a shaft sleeve and mesh into a soft metal sleeve fitted to the casing. The rotating teeth cut grooves in the softer metal and leakage is minimal.

The leakage across a labyrinth depends on the differential pressure across the seal, the number of teeth and the clearance between the teeth and the shaft. Maximum clearances are generally about 0·5 mm. Seals operating at pressure differentials below 3 barg use 3 to 6 teeth. As many as 20 teeth are needed for higher Ps.

Figure 4-13


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Figure 4-14 illustrates the use of a ‘buffer’ gas system. A neutral or clean gas is injected into Chamber ‘B’ at a higher pressure than that in Chamber ‘A’. Chamber ‘A’ is at a lower pressure than the high-pressure end because of the labyrinth seal. The combined buffer gas and high-pressure gas that leak into Chamber ‘A’, leaves through a port to a vent or disposal line. This combined gas is called sour or contaminated gas.

Figure 4-14

Labyrinth seals are almost always used for interstage sealing between impellers. They also seal at both suction and discharge ends of the compressor shaft. Most are machined from soft metals such as bronze end aluminium. If excess shaft movement occurs, the soft inexpensive metal will wear away preventing damage to the shaft.

4.6.2 Restrictive carbon rings

The restrictive carbon ring seal uses 2 or 3 piece, segmented rings held closely to the shaft by a garter spring or compressive ring band. Carbon is the usual material used because it is relatively soft and does not wear the shaft. Leakage is small across this type of seal.

The number of rings used depends on the gas pressure inside the compressor. Each seal ring is contained in a cup and a number of them are bolted together to form a cage. The cage makes it easy to remove the rings from the unit to repair or replace worn or broken segments.


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Figure 4-15


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4.6.3 Mechanical Face Seal

Mechanical seals rely on continuous face-to-face contact between machined rotating and stationary surfaces to prevent leakage of compressed gas. The seal faces are perpendicular to the shaft.

To reduce wear between the two sealing surfaces, a carbon wear ring is usually installed. The floating, unattached carbon ring is free to rotate between the stationary and rotating faces. Springs are used to force the faces together.

The sealing faces, one rotating and one stationary, are continuously pushing against each other to prevent gas from leaking between them. They must have some lubrication to prevent heat build-up and wear. Since gas is not a lubricant, oil is injected into the seal at a pressure above the adjacent internal casing pressure.

Clean, flowing pressurised oil gradually forces its way past the carbon rings faces, past the stationary seat and out of the contaminated oil port for disposal.

Gas leakage through the labyrinth seal also enters the contaminated oil port. The contaminated oil is recovered and re-used if the gas does not affect its lubricating properties.


Figure 4-16

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4.6.4 Liquid Film Seal

The liquid film is used in high pressure heavy-duty compressors. It is similar to a mechanical seal except that the seal faces are purposely separated by a measured gap. Seal oil circulates through the seal at a higher pressure than the gas and a small quantity is forced through the gap between the seal rings and shaft.

Figure 4-17

The seal oil pressure is slightly above suction pressure, so it flows into the seal gas chamber rather than leak past the seal.

Nearly all large compressors require two main seals, one on the suction side and one on the discharge side. To simplify the system, both the suction and discharge seals are identical. This is made possible by piping the discharge side back to the suction. With the aid of an extra set of labyrinth seals placed in between the last impeller and the discharge seal itself, the gas leakage is passed back to the suction. This ensures that the seal pressures remain the same. The pipe connecting the two seals is often called the balance or reference line.

Oil supplies for the seals is normally taken from the lube oil reservoir but it has its own pump and control system. Oil is supplied to the seals at a pressure slightly higher than the compressor’s suction pressure and can be controlled by a differential pressure controller or an overhead ‘Header Tank’. Seal oil that has been contaminated with hydrocarbons is de-gassed before it is returned to the system.


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4.6.5 Dry Gas Seals

Dry Gas Seals are now widely accepted and very common in the oil and gas industry. Their major advantage is that they eliminate the need for complex seal oil equipment such as pumps, cooler/heaters, degassing tanks, filters and control systems. This considerably reduces capital and running costs.

The seal is positioned perpendicular to the shaft in a similar fashion to the mechanical seal. The principle components are a mating ring (rotating) and a primary ring (stationary). The mating ring is fixed to and rotates with the shaft. Typically, it is made from very hard material such as tungsten carbide. One side of the ring has spiral grooves machined on the outer circumference of the face to a depth of 100 - 400 microns.

The primary ring is made from carbon graphite with a flat surface and is the stationary part of the seal fixed to the housing.

Figure 4-18

Both the mating and primary rings are machined and lapped to achieve a leak free seal when the shaft is not rotating. In this situation, the rings are pressed together by a spring.

As the shaft begins to rotate, the spiral grooves in the mating ring cause a pressure rise, which lifts the faces slightly apart. Because the pressure rise is dependent on the shaft speed and the seal differential, the distance between the rings is self- adjusting. The rings should only come into contact with each other on ‘start-up’ and ‘shutdown’, so wear of the rings is extremely small.

A seal gas supply is required to the outer edge of the mating disc. This gas can be supplied by the compressor itself or from an external source such as fuel gas or nitrogen generation.


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Most oilfield compressors are fitted with Tandem Seals i.e. two complete seals in one housing. The first seal isolates the process while the second seal is purely passive. If the first seal should fail, the second seal carries out the sealing function.

Figure 4-19


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4.7 Diaphragms

Diaphragms are metal components that physically divide the individual stages (impellers) in a multi-stage compressor.

Effectively, they are individual casings for each impeller and allow the gas to flow from the discharge of one impeller to the suction of the next. In some large, horizontally split compressors, the heat generated by compression causes diaphragm expansion problems. Circulating cooling water to the diaphragms can prevent this.

Figure 4-20

The diaphragm’s material of construction is important, Cast iron is good for most situations, but cast steel, aluminium cast bronze, or special alloys may perform better. Cast iron diaphragms cannot withstand high differential pressure and may break under abnormal conditions.


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4.8 Guide Vanes

Compressor performance is affected by the velocity and direction of the gas as it enters the eye of the impeller. The compressor capacity and gas pressure rise are increased slightly if the gas enters the impeller in a direction opposite to that of the rotation of the impeller.

The compressor efficiency is higher if the gas enters the impeller in the same direction as the rotation of the impeller.

Guide vanes located in the diaphragms direct the flow path of the gas to the next impeller.

The first impeller in a multistage unit may have vanes that are adjustable. They can be moved manually or automatically via linkages located outside the compressor casing to alter the flow direction of the gas.

Figure 4-21

They can also be used to control the capacity of a compressor. Capacity is reduced by moving the vanes so they restrict the flow and change the direction of gas into the first stage impeller.


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4.9 Couplings

Couplings transmit rotation from the driver shaft to the compressor shaft. If a gearbox is situated between the driver and the compressor, a coupling attaches the driver shaft to the inlet gearbox shaft and another coupling attaches the outlet gearbox shaft to the compressor shaft.

The couplings must be able to withstand the shock of a sudden change in compressor load or startup and shutdown of the driver. They must be flexible enough to transmit power from the driver to the compressor at high speed when the two shafts are not perfectly aligned.

In some instances it is almost impossible to perfectly align the two shafts because of operating and load variations, the temperature difference of the driver and compressor result in one expanding slightly more than other. The coupling must be able to accommodate and overcome the misalignment.

In each type, the driver shaft attaches to one half or hub of the coupling and the compressor shaft attaches to the other. The shafts are usually keyed to the couplings’ hubs; a typical coupling is shown in Figure 4-22.

Figure 4-22


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5.0 COMPRESSOR LUBRICATION

Small, low-pressure, compressors can be lubricated by a variety of lubrication systems. These include grease cups, top or side mounted oil drip glass reservoirs or crankcase pumps similar to those used in automobiles. Larger units normally have a separate external circulation system.

Compressor parts fit together with very little clearance. Bearings are used to support the weight of the shaft and impellers and minimise radial movement. The shaft must be free to rotate, but not vibrate. Without a lubricant, the bearing surfaces would wear out quickly.

Oil is supplied to bearings through machined holes and grooves. The oil forms a film between the rotating shaft and the bearing that separates them. This reduces friction and heat but some heat is still generated and the lubricating oil must carry it away. Before the oil is reused, it must be cooled, cleaned and re-pressurized.

Oil is pumped from the reservoir to a filter, which removes solid particles. Clean oil then flows under pressure to each bearing where it removes the heat of friction. It then flows to a cooler to remove the heat collected in the bearings and then back to the reservoir for re-circulation (see Figure 5-1).

Figure 5-1

This simple system works for small machines that can be shutdown for filter cleaning and repairs. The large machines used in the oil industry are often required


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to operate for long periods of time, to achieve this, a more complex system is required.

Machine failure occurs very quickly if lube oil flow and pressure are not maintained. The metal parts come into contact when there is no oil film to separate them. The friction of fast moving metal-to-metal contact quickly builds up sufficient heat to melt bearings and damage the shafts.

To ensure continuous operation of any lubrication system, spare pumps and filters must be ready for immediate use when the operating units fail (see Figure 5-2). An auxiliary lube oil pump (with a different power source from the primary pump) is included on larger installations and starts automatically if the primary pump fails. The auxiliary pump should be checked regularly to ensure its operation when needed.

Duplex filters with switching and isolating valves are usually provided to allow filter changes without compressor shutdown. Most duplex filter systems are equipped with interlocking three way valves to ensure that one filter is always in service. A differential pressure gauge is installed to indicate the condition of the filter in service. An increase in the gauge reading indicates that the in-service filter is becoming blocked and the elements should be replaced.

The discharge pressure of an oil pump is determined by the bearing system of the compressor that it is supplying. A normal oil supply pressure recommended by the manufacturer is enough to force oil into bearing surfaces at the proper rates. Pressure regulators may be used to reduce oil pressure for lubrication points requiring less pressure.

The type of lube oil system used depends on the size of the compressor and the type of driver. Machines up to 1000 hp often have a pump built into the casing that is gear-driven from the compressor shaft and has an external cooler and filter. This is known as an integral or internal lube oil system. Larger machines usually have external oil systems.

For a compressor driven by a motor or reciprocating engine, an external lube system is provided, serving only the compressor and possibly a gearbox and coupling between the compressor and the driver.

When a gas turbine is the compressor driver a single lube oil system can in some instances supply both the turbine and the compressor.


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Figure 5-2


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5.1 Lube Oil System Flow Description (See Figure 5-2).

Oil pumped from the tank flows through a cooler, where the heat generated in the compressor and turbine bearings is removed. After the cooler, a slipstream of oil is returned to the oil tank to maintain a constant oil pressure to the turbine bearings.

From the cooler, the oil flows through a filter to remove solid particles. From the filter, the oil stream divides into two flows. One stream flows to the turbine bearings and the second stream, after further pressure reduction, flows to the compressor bearings. The bearing outlet oil lines combine and flow back to the tank.

A back-up or pressure reserve system provides short term lubrication only. The back-up unit is a tank partially filled with oil under normal operating pressure. Additional reserve pressure is provided by an internally mounted, inflated bladder of nitrogen or other inert gas. This back up system is sized to minimise the normal pressure fluctuations that occur during routine or automatic pump changeovers, which could generate a compressor trip.

When the oil pressure drops, the bladder gas expands, forcing oil into the lube system for a short period. The back-up system refills when the regular oil pump is back on-line and the bladder gas is compressed.

The system shown has three oil pumps:

The pre-lube pump is used at start-up to pre-lubricate the bearings. It stays on until the compressor is started and reaches operating speed.

When the turbine reaches operating speed, the main oil pump automatically starts and the pre-lube pump shuts down when operating pressure is achieved.

The auxiliary pump is normally powered by d.c. current or batteries. It starts automatically when the main oil pump shuts down and no a.c. power is available. It supplies lubrication only during rundown and after the unit stops.

Sight glasses are often placed in the bearing oil outlet lines so that the operator can actually see that there is sufficient oil flow to individual bearings.

Instrumentation applied to any lube oil system is dependent on a number of factors. These include: system-operating pressure, nature of the gas being compressed and the type of driver.

The following instrumentation is included in most medium to high pressure systems:


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5.2 Lube Oil System Instrumentation

An oil heating element in the lube oil storage tank (normally electric but can be heating coils using a circulating heating medium), is switched on by a thermostat to keep the lubricant – at the required temperature.

Low level switches in the oil reservoir initiate alarms and/or pump shutdowns.

Oil pumps are oversized to ensure that there is always sufficient flow. Discharge oil pressure is measured and controlled by returning a slipstream of oil to the lube oil storage tank through a recycle valve.

Differential pressure instruments compare the pressure upstream and downstream of the oil filters. The signals are sent to indicators, alarms and shutdown devices. Sometimes the P across a filter has to be determined by reference to inlet and outlet pressure gauges.

An oil temperature controller regulates cooling to maintain the correct oil temperature to compressor bearings.

Auxiliary pump instruments such as a power failure switch or low oil pressure switch sense failure of the main oil pump and immediately start the second and/or third auxiliary pump to maintain compressor lubrication.

Thermometers and pressure gauges monitored by operations personnel show that the heater element, pumps and cooler are working properly. The temperatures of compressor bearings and gears are also monitored.

Shutdowns or trips are provided on larger units to shutdown the entire unit i.e. compressor, driver, and auxiliary equipment.

The following lube oil conditions may initiate a trip:

Low-low oil reservoir level,

Low-low oil pressure,

High-high/low-low oil temperature,

High-high differential pressure across the filter.

A lube oil system should continue to operate even when the compressor is shutdown. Compressor rotation takes some time to stop, especially on large units and oil lubrication is essential during this rundown period. Lube oil systems should continue to operate at least until the compressor has completely stopped rotating. Even after rotation has stopped, it is desirable to continue oil circulation until the bearing, seals and ancillary equipment have cooled evenly. Also, when oil circulation is continued, the compressor can be restarted quickly.

Most oil systems require a ‘pre-lube’ period prior to start-up. The oil pump is started and oil flows through the compressor bearings for a period of at least 5 minutes before continuing the compressor start-up procedures. This action assures lubrication during the start-up, when surging and other harmful events can occur.

To summarise, full pressure lubrication to a compressor is required before start-up, during operation and after shutdown.


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6.0 COMPRESSOR OPERATING CURVES

Each centrifugal compressor has operating curves similar to those shown in Figure 6-1. These show the relationship between capacity and compression ratio at various compressor speeds. Compression ratio is normally used but some manufactures will use differential pressure and both of these can be expressed as a percentage, as can capacity. The operating limits of the compressor are fixed by a surge condition on the left and a maximum capacity limitation (stonewall) on the right.

Assume only 80% of the design gas flow rate is available to the compressor. At this flow rate and 100% speed; the compression ratio is going to be 108% of design. In other words, the compressor will not operate at 100% compression ratio at 100% speed and 80% of capacity.

At lower capacity, the compression ratio must increase. Remember, that the compression ratio is the absolute discharge pressure divided by the absolute suction pressure. When the compression ratio rises, the discharge pressure rises or the suction pressure goes down, or both occur.

Figure 6-1


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The only way the compressor can be operated at 100% compression ratio and 80% capacity is to slow the unit down to about 97% speed. Thus, speed control is necessary in order to hold a constant compression ratio at different gas flow rates or vice versa. The method for changing speed will depend upon the type of driver, which may be connected to the compressor via a gearbox.

Common types of variable speed drivers are:

Gas Turbines.

Direct Current Electrical Motors.

Internal Combustion Engines.

Steam Turbines

If the compressor speed cannot be varied, the 100% speed curve will be the only operating curve for the unit.

It is important to understand that at a given compressor speed and gas flow rate, the compressor is going to develop a certain compression ratio, whether it is required or not. The best way to change an operating point is to change the speed.

Another point shown on this set of performance curves is that this compressor will not develop 100% compression ratio at a speed below 95% of design. If it runs at 110% speed, it will always exceed 100% compression ratio, regardless of the gas flow rate. These speed ranges do not apply to all compressors, but they are typical.

If the gas flow to the compressor is less than 58% of the capacity at 100% speed, some discharge gas will have to be recycled to the suction in order to prevent the compressor from surging.

Each compressor unit has an operating curve that is supplied by the manufacturer. The curves may be different from those shown, but they will all have the same general shape and will all be limited by a surge line on the left and a stonewall condition on the right.


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Worked Example: How to Read a Curve

When the compressor speed is 95%, the gas flow rate is 80 MMscf/d and the suction pressure is 203 psig, what is the discharge pressure? See Figure 6-2

Gas Flow rate = 80 MMscf/d

Speed = 95%

Compression Ratio (from curve) = 3·9

Absolute Suction Pressure = 203 + 15 = 218 psia.

Compression Ratio (CR) = Absolute Discharge Pressure

Absolute Suction Pressure

Re-arrange the equation to determine absolute discharge pressure:

Absolute Discharge Pressure = CR x Absolute Suction Pressure

= 3·9 x 218 = 850 psia


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Figure 6-2


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6.1 Compressor Control

A compressor is designed to compress a fixed volume of gas from a designated suction pressure to a designated discharge pressure. In operating situations, the suction pressure, discharge pressure and gas flow rate are seldom at design conditions and can fluctuate.

The function of the control system is to safely operate the machine at the lowest cost within safe operating limits, while keeping operating supervision to a minimum.

Centrifugal compression controls serve three purposes:

Capacity control

Surge prevention

Equipment protection.

Definition of Capacity

The volume of gas that a compressor handles depends on the size of the impellers, the speed at which they rotate and the compression ratio they develop. The impellers in an operating compressor running at a constant speed and compression ratio will process a certain volume of gas.

This volume is the same at any suction temperature or pressure.

Gas flow meters automatically correct readings so that they give the flow rate at standard conditions instead of at operating conditions. Most meters will still have a factor by which the reading must be multiplied to obtain the correct flow rate. The factor is necessary for the instrument’s range.

Because gas is rarely at atmospheric pressure and 15C (the pressure is usually much higher) the volume flowing at operating pressure is only a fraction of its theoretical volume at atmospheric pressure. For example, 1,000 volumes of gas at atmospheric pressure would occupy only 20 volumes at a gauge pressure of 50 barg.

The volume at operating pressure is 20 units, but if the pressure was lowered to atmospheric, the gas would expand to 1,000 units, which is what the flow meter reads.

Assume the capacity of a compressor operating at a certain speed and compression ratio is 350 cf/min. It will pull in this volume of gas at atmospheric pressure, under a vacuum or at high pressure. The flow meter volume of gas will depend upon its pressure and temperature.

The operating curves supplied by a compressor manufacturer sometimes show the capacity at operating conditions in units of cubic metres/min. In order to find the capacity in standard cubic feet/day, the volume at suction temperature and pressure must be converted. To convert cubic metres to cubic feet, divide the cubic metre by 0·02832. To convert the volume at operating conditions to standard conditions the following formula is used:

P1V1 = P2V2


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T1 T2


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Since standard conditions are always the same, a factor derived from these constants can be calculated.

i.e. 14.7 psia/ 1 bara (standard pressure)

520o R for Fahrenheit scale/ 288 K for Celsius scale (standard temp)

0.0283 (multiplication factor to convert cubic feet to cubic meters)

35.315 (multiplication factor to convert cubic meters to cubic feet)

Examples

The capacity of a compressor at 100% speed and 100% compression ratio is 1770 cf/min. Gas enters at a pressure of 363 psig and a temperature of 86 F. Calculate the capacity in standard volume units.

Standard Volume Equation = P1V1 = P2V2

T1 T2

V2 = P1V1 x T2

T1 P2

Where:

V2 is standard volume

Actual Volume per day = 1770 x 1440 = 2.55 MMcf/d

Suction pressure, P1 = 377.7 psia

Suction temperature, T1 = 546o·R

Substitute in equation = 377.7 x 2.55 x 520

546 14.7

= 1.76 x 35.37

Standard volume per day = 62.4 MMscf/d


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Suction pressure in the above compressor drops to 290 psig and the temperature is unchanged. What standard volume will the compressor handle?

Actual Volume per day = 2.55 MMcf/d

Suction pressure, P = 304.7 psia

Suction temperature, T = 546o R

Substitute in equation = 304.7 x 2.55 x 520

546 14.7

= 1.42 x 35.37

Standard volume = 50.3 MMscf/d

In the examples above, the real volume of gas flowing to the compressor was the same, but the standard volume of flow metered volume changed 20% when the pressure changed 20%. Compressor operators are concerned with standard gas volumes. From the above examples, it can be seen that the standard volume of gas that a compressor will handle varies proportionally with suction pressure.

If the suction pressure drops by 5%, the compressor will handle 5% less gas at the same speed and compression ratio.


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6.2 Capacity Control

This is of particular importance in a fixed speed machine. If the unit is to handle more gas at the same compression ratio, the suction pressure will have to be raised. In most compressor operating situations, the suction pressure does not vary more than 10%.

If the capacity is based on an average suction pressure, the maximum change in capacity with change in pressure will be ± 5%, which is about the accuracy of the operating curves. When considering capacity, it is assumed that suction pressure changes vary no more than ± 5%, and that no corrections are made for them. However, if the suction pressure drops 10%, the capacity of the compressor in standard volume will be 10% less, and vice versa.

If the suction pressure and gas flow rate to a compressor were to remain constant, control systems would not be needed. However, operating conditions vary and a control system provides the means to operate the unit when conditions change.

The varying process conditions are:

Suction temperature, pressure and flow rate

Discharge temperature, pressure and flow rate

Compression ratio

Gas density

The above variables combine in a compressor to form its capacity characteristics, which are defined in the operating curves.

At any compressor speed, the compression ratios developed at various gas flow rates (compressor capabilities) are defined by the curve for that speed. The unit will operate only on a single curve at one speed.

For example (refer to Figure 6-3), at a gas flow rate of 84% of design and a speed of 100%, the compression ratio this machine develops will be 3·2. Design compression ratio at 100% flow is 2.9. At 84% flow, the discharge pressure will be higher than design or suction pressure will be lower than design.

But, suppose suction and discharge pressure are at design conditions, which means that the required compression ratio should also be at 100% design, which is 2.9.

If a compressor is going to deliver gas at compression ration of 3·2 at 84% flow, the control system must remove the extra 0.3 of compression ratio that the unit develops at 100% speed and 84% gas flow.


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Figure 6-3

Assume that the gas flow is 84% of capacity, speed is 100% and suction and discharge pressures are at their design points, which requires a compression ratio


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of 3·0. The compression ratio developed at 84% flow and 100% speed is 3·2, but operating conditions require a compression ratio of only 3·0.

The capacity control system must get rid of the excess compression ratio caused by high discharge pressure. This can be achieved using a throttling valve in the discharge line.

Assuming that the design suction pressure is 145 psig and the discharge is 465 psig, the pressure drop across the throttling valve is determined as follows:

C.R. at 84% flow, 100% speed = 3·2

Suction pressure (gauge) = 145 psig

Suction pressure (absolute) = 145 + 15 = 160 psia

Absolute discharge pressure at

3·2 compression ratio = 160 x 3·2 = 512 psia

Discharge pressure, gauge = 512 – 15 = 497 psig

Design discharge pressure = 465 psig

Excess discharge pressure = 497 – 465 = 32 psig

Figure 6-4

From the above calculations, it can be seen that the compressor discharge pressure is 32 psig higher than the design. This excess can be removed with a capacity control valve located in the discharge line. The capacity control valve is opened until the discharge pressure gauge reads design, at which point CR is 3.0 and capacity is 96% of the design pressure. This is illustrated in Figure 6-4.


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Another means of reducing excess compression ratio at reduced capacity and 100% speed is by throttling the suction gas entering the compressor. In this example high CR is caused by low suction pressure (see Figure 6-5).

Using the same design pressures as before, the calculation for determining the amount of pressure drop across the capacity control valve in the suction line is as follows:

CR at 84% and 100% speed = 3·2

Discharge pressure gauge = 465 psig

Discharge pressure absolute = 480 psia

Absolute suction press = 480 = 150 psia

3·2

Suction press gauge = 135 psig

Design suction pressure = 145 psig

Excess suction pressure = 145 – 135 = 10 psig

Liquid Disposal Liquid Disposal

Figure 6-5

In this case, the suction valve is opened to increase the suction pressure from 135 psig to 145 psig, which will result in a CR of 3.0 and capacity 96% of design.


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A third method of controlling capacity of 84% gas flow and 100% speed is by recycling some discharge gas to the suction to the compressor (see Figure 6-6).

Figure 6-6

Gas leaving the discharge of the compressor is hot and if it were recycled, it could result in overheating and damage to the compressor. Therefore, the recycle gas line is branched from the outlet of the discharge separator, where it has been cooled and any resultant condensate removed.

The volume of recycle in the example is 16% of the total capacity of the compressor. In this situation, flow to the compressor is 100% of design of which 84% flows out of the discharge line to the process and 16% is recycled back through the compressor. The compressor is using 100% power for only 84% of flow, some of the driver power is wasted.

In some units, adjustable guide vanes are installed at the inlet to the first impeller. Although these vanes throttle the suction gas flow, their main purpose is to control the rotation the gas into the impeller eye. Rotation control may be in the same direction as, or opposite to, impeller rotation. This action either loads or unloads the impeller, which results in changing the capacity of the machine.

These guide vanes can be manually adjusted, or they can be automatically adjusted with a suction pressure controller. Guide vane rotation control is a more efficient method of capacity control than suction or discharge throttling. Its disadvantage is initial cost and inaccessibility for maintenance. The flow control range is usually good with this method.


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Guide vanes are justified if the compressor operates at reduced loads for long periods. The load on the driver with guide vane control at reduced capacity will be slightly more than the normal power required (see Figure 6-7).

Liquid Disposal Liquid

Disposal

Figure 6-7

The four capacity control systems described; throttling suction gas, throttling discharge gas, recycle gas and guide vanes can be used on machines operating at a constant speed. A single unit usually has only one of these control systems installed but occasionally recycle and guide vane systems are installed together.

If the unit has a variable speed driver, the primary capacity control is by varying the speed of the machine (refer to the operating curve). At 84% flow rate, the machine will develop the design compression ratio of 3·0 if the speed is reduced to about 97%. A throttling valve in the suction or discharge line is a secondary control used for maintaining the desired compression ratio when speed alone will not provide the necessary capacity control.

Speed control is the simplest and most economical means of controlling the capacity of a compressor because the other control methods result in a waste of driver energy. This waste is unavoidable in a constant speed machine. The recycle control results in the greatest energy consumption by the driver, so it is the least preferred method of control. Throttling the suction or discharge gas to maintain capacity control at reduced gas flow rates, consumes less driver energy than recycling gas.

The operating curves supplied by compressor manufacturers may have different units from those shown. For example, the capacity scale may be shown in actual volume at suction temperature and pressure MMcf/d; or the units may be in standard volume units, MMscf/d.


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The vertical scale on the left of the compressor operating curve may be expressed as:

differential pressure (discharge pressure – suction pressure),

head (height) of gas at the discharge side,

discharge pressure at a constant suction pressure,

as a percentage of any of the above.

Any combination of these units has some limitations.

Operating curves with discharge pressure as the left-hand vertical scale apply only to gas at constant density, temperature and suction pressure. A different set of curves is required for each different density, temperature and suction pressure.

Normally, the compression ratio curve is used because the compression ratio developed by a centrifugal compressor is not affected by normal changes in the temperature or pressure of the gas flowing to it. In other words, a centrifugal compressor operating at a certain speed and percentage capacity will develop the compression ratio shown on the curves for gas of a constant composition at any suction pressure and within a temperature range of ± 20F.

Summary of Capacity Control

The capacity of a compressor must be controlled so that it is the same as the volume of gas flowing to it.

Capacity is affected by speed and compression ratio.

Capacity is controlled on variable speed machines by adjusting speed and compression ratio.

Capacity is controlled on constant speed machines by adjusting the compression ratio.

This is accomplished with:

a control valve in the suction line

a control valve in the discharge line

a control valve in the recycle line

guide vanes in the compressor inlet.


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6.2.1 Capacity Control of Single Compressor Units

Constant Speed Compressors

A constant speed unit has only one operating curve, which is usually the 100% speed curve. Capacity is normally changed by adjusting the compression ratio or by recycling discharge gas to the suction side.

There are two common methods of changing the compression ratio:

Throttle the discharge gas,

Throttle the suction gas.

Both control methods, plus recycle control were discussed in the previous section but did not include the method by which the capacity control valve in the suction, discharge or recycle line is opened or closed.

One method of regulating the position of a capacity control valve is via a pressure controller in the compressor suction line. The set point on the suction pressure controller is positioned by the operator and will vary depending on process conditions.

The controller is connected to the suction line and it continually senses suction pressure. If there is a reduction in gas flow rate to the compressor, suction pressure will fall. The pressure controller will sense the drop in pressure and send a signal to the control valve in the discharge to throttle, which in turn will raise the compression ratio and lower the capacity of the compressor. A ratio-limiting device is used in this type of control to prevent the compression ratio from exceeding the capability of the compressor.

If gas flow to the compressor rises, suction pressure will rise and the suction pressure controller will signal the discharge control valve to open, which will lower the discharge pressure and compression ratio and increase the capacity of the compressor.

A suction pressure controller can also be used to regulate a control valve located in the suction line. Action of the pressure controller is the same as in the preceding system. When the flow of gas rises, suction pressure will increase and the pressure controller will signal the control valve to close slightly.

A suction pressure controller can also be used to regulate flow in recycle system. If the gas flow-rate drops, suction pressure will drop and the pressure controller will signal the valve in the recycle line to open, which will increase the gas flow to the compressor.

The final method of controlling the capacity of a constant speed machine is by regulating the position of guide vanes. A suction pressure controller is commonly used to adjust the guide vanes to hold a constant suction pressure. It the gas flow rate drops the pressure controller will rotate the vanes to reduce the capacity of the compressor or vice versa.

Variable Speed Capacity Control

Speed is usually controlled to hold a constant suction pressure. If flow to the machine increases, suction pressure will rise and the suction pressure controller will


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signal the speed controller to increase. A reduction in gas flow rate will result in a decrease in suction pressure and the suction pressure controller will signal the driver to lower its speed. Speed changes must stay within the compressor speed range.

The speed controller can also be regulated with a pressure controller in the discharge line set to hold a constant discharge pressure or a flow controller can regulate the speed to maintain a constant gas flow to the unit at varying suction and/or discharge pressures.


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In each of the capacity control systems described, the control point was the suction pressure to the compressor. The main reason that suction pressure is used as the control point is that a small change has a greater effect on compression ratio than the same change in the discharge pressure.

For example, with the compressor discussed in the previous examples, a change in the suction pressure of 15 psig (10% change) at a constant discharge pressure will change the compression ratio approximately 10%. However, a change in discharge pressure of 15 psig at constant suction pressure, will change the compression ratio by only 3%.

The capacity of a compressor is determined by the speed and compression ratio and suction pressure is the usual control point.

There are two exceptions to suction pressure control:

When a constant discharge pressure is required, a pressure controller in the discharge line regulates the speed, control valve or guide vanes.

When a constant gas flow rate is required, unit speed, a control valve or guide vanes are regulated via a flow controller.

Speed Control Limitation

Using speed to control capacity has one limitation i.e. it can only be reduced so far and still develop the compression ratio required. For a machine with the operating curves as shown in Figure , the minimum speed that will provide a compression ratio of 3·0 is 95%. This is the lowest speed at which the machine can operate when a compression ratio of 3·0 is required.

Figure 6-8

If the gas flow rate to the unit is 50% and a compression ratio of 3·0 is required. The speed is reduced to 95%, but the unit will surge if the flow drops below 54%. As will be seen later, the surge control system will recycle enough gas to make up the difference between actual flow rate (50%) and surge flow rate (54%).


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The main point is that there is a limit to the amount of speed reduction. Below that limit, the compression ratio is affected and gas recycle is required. This is not good operation when two units operate in parallel at low gas flow rates. At low gas rates, it is difficult to adjust the speed of two units (manually or automatically) so that each has the same capacity and compression ratio.

Quite often, the load will shift from one unit to the other and then reverse to the first one. The unit that is loaded will start to surge. This situation is corrected by increasing the recycle to both machines, either manually or by raising the setting in the surge control system.

Figure 6-9


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6.3 Surge Control

The operating range of a centrifugal compressor is restricted by a surge condition at low flow rates and a stonewall condition at high flow rates. The surge limitation is the more critical of the two.

Inside the compressor there are no check valves or other mechanical devices to prevent discharge gas from flowing backwards towards the suction side of the unit. Back-flow can occur when a combination of conditions is met i.e. low gas flow rate, high compression ratio. A high compression ratio can result from a high discharge pressure, low suction pressure or a combination of the two.

Discharge pressure can be represented by head. Figure 6-10 below shows the sequence of events during a surge.

Figure 6-10

Surge is a spasmodic reversal of flow backwards through a compressor and is caused by the flow being below the compressor’s minimum stable level.

The operating curves show the combinations of flows and compression ratio at various speeds that result in a surge condition. For example, in Figure 6-1, at 100% speed, surge will occur at a flow rate slightly above 58% and a compression ratio slightly over 110%. At these conditions of low flow and high compression ratio, discharge gas will flow back through the compressor.

As soon as backflow occurs, the discharge pressure drops and forward flow starts again. When discharge pressure rises to the surge point, flow reversal happens again.


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Surge may happen several times a minute, or several times a second, depending upon the compressor size, speed etc.

Surge is made obvious by the definite sounds produced. These may be a whine, scream, moan or a sharp tapping noise, accompanied by high vibration, fluctuating flows and pressures. A rise in the compressor discharge temperature is also a possibility.

Surge may cause severe damage to the compressor and piping. Check valve flappers can break and internal damage to the compressor may be caused. In addition, the back flowing gas tries to reverse the rotation of the machine, so added strain is put on couplings between the compressor and driver.

Surge results from a combination of low flow and high compression ratio. It can be avoided by maintaining the flow above the surge point. This is achieved by recycling some discharge gas to the suction side to maintain enough flow to the unit so that the surge condition is avoided.

Surge can also happen during start up and shutdown of a centrifugal compressor. Surge protection must be provided during these operations, particularly on large machines.

Surge is avoided by recycling enough gas from the discharge side to the suction side so that the flow through the compressor is above the surge point. Since discharge gas is hot, it must be cooled before it is recycled to the suction side of the machine. The following drawing shows a recycle system in a unit with a cooler on the discharge gas. The control valve in the recycle line is called an anti-surge or recycle gas control valve.

Figure 6-11

The anti-surge gas control valve is regulated by a flow controller (Surge Control), which measures gas entering the compressor. The set point of the flow controller is set above the flow rate at which surge occurs.

If the inlet flow to the compressor approaches the set point of the controller, the anti-surge valve opens and allows enough re-cycle gas to flow into the suction line so that the total flow to the compressor is above its surge point.


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For example, if a compressor with the operating curves shown in Figure 6-9 is operating at 100% speed, the flow controller would need to be set at a point above 56% flow to prevent surge.

The controller would probably be set about 62% to allow a margin of safety. If the inlet flow to the compressor is less than 62%, the flow controller signals the anti-surge control valve to open sufficiently to bring the total flow to the compressor to 62% of the design rate.

On variable speed machines, the gas flow rate that results in surging varies with speed. The set point in the surge flow controller has to be reset each time the speed changes. This can be done automatically with a speed controller, which resets the set point on the surge flow controller when speed changes. Normally, the surge flow controller set point is reset by compression ratio changes.

As previously mentioned, surge can occur during start-up and shutdown. In fact, these are the times when surge is most likely to occur. It can be prevented at these times by fully opening the anti-surge control valve. This equalizes the suction and discharge pressures and prevents a high compression ratio from occurring.

One way of controlling capacity on a constant speed compressor is to regulate the flow of recycle gas. Although identical in operation, this is not an anti-surge control system.

Surge control prevents flow from dropping to the surge condition. Recycle control regulates flow (capacity) to maintain a certain compression ratio. Machines using a recycle type of capacity control could also have a separate surge control system.

Each compressor in a multi-stage compressor system may have its own surge control system. The usual exception is that of compressors operating in series. In this case, the recycle gas flows from the outlet of the last stage discharge cooler to the inlet to the 1st stage suction scrubber.

Summary of Surge Control System

Surge occurs at a low flow rate and high differential pressure in a centrifugal compressor. At these conditions, gas back-flows through the compressor unit.

The primary means of preventing surge is to prevent flow from dropping to the point that surge will occur by recycling discharge gas to the suction.

Secondary control devices that measure compression ratios in some form are often used to override the primary flow control device and increase flow of recycle gas to prevent the compression ratio from reaching the surge condition.

The surge control system should be set to avoid surging condition at the lowest power consumption of the driver.


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6.4 Stonewall

As mentioned previously, the minimum flow to a centrifugal compressor is the flow when it enters a surge condition. The maximum flow at a given speed is fixed by the stonewall limitation. Stonewall occurs at high flow rates when the impellers raise the velocity of gas to the point that it reaches sonic speed i.e. the speed of sound.

At sonic speed, no more gas can be forced through the compressor, regardless of any changes that are made. In other words, the unit has hit a stone wall. When this occurs, shock waves cause a rapid decrease in pressure. Consequently, the unit is kept away from the operating range near the stonewall or high flow point.

The stonewall effect is of minimal concern in most oilfield and gas processing compressor applications where the gas is methane, ethane, propane or a mixture of the three. The conditions of the flow, speed and compression ratio that result in stonewall seldom occur in most oilfield operations. Consequently, controls to prevent stonewall are rarely used.

6.5 Gas Density

The compression ratio developed by a compressor depends on a number of factors.

One is the density or weight of the gas. Natural gas is about two thirds as heavy as air, i.e. air is 50% heavier than gas. If air entered the compressor instead of gas at the same suction temperature and pressure, the compression ratio would be 50% higher than with natural gas. The driver would also require 50% more power to compress the air.

In most compressor applications, there is not enough change in gas density to effect compression ratio or driver power requirement. In a small percentage of cases where the gas density changes over a wide range, a control system is often included to maintain a constant density of gas entering the compressor.

The system provides for introduction of a lighter or heavier gas, depending upon the requirement, to maintain a constant density of gas entering the compressor. Some compressor manufacturers call this type of control a ‘molecular weight control’; others call it ‘relative density’ or ‘specific gravity control’. Molecular weight, density, relative density and specific gravity of gas are all related. A percentage change in one will have the same percentage change in the others.

The main reason for having a density control system is to prevent overloading the driver when the gas density increases. A secondary reason is that the capacity control systems must be reset each time the density changes.

Capacity control systems use speed or throttling valve adjustments so that the compressor operates at a compression ratio on the operating curve for that unit at the gas flow rate. A higher density gas will have a performance curve parallel and higher than the normal curve. The capacity control will have to be set for the higher curve when gas density increases. The density control system does not replace a capacity control system. It controls the quality of gas entering the machine. The capacity control system operates in its normal way.


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6.6 Critical Speed

All solid materials have certain vibration levels called natural frequencies. A glass for example, will shatter from sound vibrations that match its natural frequency.

If a material is subjected to vibration at its natural frequency for any extended period of the time, it will usually disintegrate.

One important part affected by vibration in a centrifugal compressor is the rotating element (shaft and impellers). The rotor has a natural frequency of vibration in the same way as the glass. This frequency occurs at a certain speed of rotation, which is called the critical speed.

If the unit operates at its critical speed for an extended period, some form of breakage is almost certain to happen. In almost all cases, the critical speed is below the normal operating speed range of a compressor. In fact, the minimum operating speed is usually specified by the compressor manufacturer and is safely above the critical speed.

Secondary or harmonic natural frequencies of lesser size occur at higher speeds than the low speed or first critical. The speeds that result in natural frequency vibration are called the first critical (lowest speed), second critical etc. Most centrifugal compressors operate at a speed between first and second critical speed.

During start up and shutdown, the compressor speed usually passes through the first critical speed. The passage through the critical speed ranges should be made as quickly as possible to avoid the possibility of damage caused by high vibration.


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7.0 COMPRESSOR MONITOR AND PROTECTION DEVICES

Monitoring instruments simply indicate what is happening in the compressor system, they have no direct control function. These include pressure gauges, thermometers, level gauges, sight glasses, flow recorders, differential indicators, vibration monitors and similar devices.

7.1 Alarm and Shutdown Protection

Alarm and shutdown devices are part of the control system that protects the equipment from damage due to mechanical failure or from exceeding process limits such as temperatures, pressures, vibrations, etc.

Alarms warn the operator that process conditions are approaching the maximum limits of their normal operating range. In most cases, quick corrective action will clear the alarm condition before it reaches a shutdown condition.

Shutdown devices usually sense the same process condition as the alarms, but they trip when that condition reaches the point that it is unsafe to continue equipment operation. They shutdown the compressor to prevent damage.

Alarm and shutdowns for most compressor systems are divided into the following categories:

Start-up/Shutdown Sequences.

Compressor Process.

Mechanical Failures.

Lube/Seal System.

If the compressor is driven by a turbine or engine, the alarm and shutdown system is often included with the driver shutdown system. Different manufacturers have different shutdown trip points.

Alarms and shutdown trips are usually indicated on some type of annunciator panel and quite often, several annunciator panels are provided. When an alarm device is activated, it lights a bulb on the panel and gives an audible sound.

The sound can be silenced by pushing a SILENCE or ACKNOWLEDGE button but this does not make the light go out. The light will only go out if the alarm condition is corrected and the variable drops below the activation setpoint. In some cases, a RESET button must be pushed to clear the light once the alarm has cleared.

When a shutdown device causes a trip, it signals the driver to shutdown and also signals the annunciator panel. As before, a bulb lights (usually red) showing the trip point and an audible signal is heard.

7.1.1 Start-up/Shutdown

Large compressors are often equipped with automatic start and stop sequence systems that are started by pushing a START or STOP button. If a step in the sequence fails, the start-up is aborted and an alarm signal indicates on the alarm panel showing the cause.

If the compressor driver is a gas turbine, the automatic start/stop sequence control panel usually includes both the driver and the compressor.


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For example, if a turbine igniter fails during start up, start up is aborted. An alarm light appears on the control panel indicating the failure.

7.1.2 Monitoring System

There is usually very little an operator can do to correct a failure during the automatic start-up sequence. In most cases, start-up will abort. The cause of the failure is usually indicated on the control panel. The operator must find and correct the cause of failure before trying to restart.

Process Alarms and Shutdown

While a compressor is operating, process alarms and trips protect the system against damage caused by process upsets.

The alarm switch is activated as a warning before the shutdown switch is activated. Two devices, one for alarm and a second for shutdown are located at these points. For example, a compressor suction scrubber has two float switches.

The first controls the level and also has a high level alarm or pre-warning setpoint built in.

The second float is mounted above the first at a point where liquid carryover might occur. This could damage the compressor, so a shutdown is initiated if the level rises to the location of the shutdown float on the scrubber.


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CAUSE AND CORRECTIVE ACTION FOR PROCESS ALARMS AND SHUTDOWNS

ALARM OR SHUTDOWN

POINT CAUSE OF TRIP CORRECTIVE ACTION

Suction Scrubber

High liquid level Manually drain liquid

Reset level controller, lower level

Suction Low Pressure Check that the suction valve is not partially closed or for other restrictions and correct.

Check for high speed on variable speed machines and reduce.

Check for low recycle gas flow in constant speed machines and increase as required.

Check for drop in discharge pressure and increase as required.

Discharge High Temperature Check for increase in suction temperature and reduce if necessary.

Check for increase in compression ratio and reduce by lowering speed or increasing flow of recycle gas.

If 1 and 2 are OK, a mechanical problem inside the compressor is indicated.

Discharge High Pressure Check discharge piping for restrictions due to partially closed valves and correct.

Check pressure drop across discharge gas cooler for hydrate formation or plugging and correct if indicated.

Check for increase in suction pressure and lower speed or reduce recycle gas.

Check the flow of cooling medium (air or water) and increase as necessary.

Check pressure drop for fouled tubes. Fouling can be scale, hydrate, corrosion


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products etc. Correct fouled condition.


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7.1.3 Mechanical Failure Alarm and Shutdown

A mechanical failure of the compressor is usually either a bearing failure, which results in high temperature at the bearings, or an internal part failure such as an impeller or diffuser, which results in vibration.

Mechanical failure alarm/shutdown parameters include high bearing temperatures and vibration. Bearing temperature elements may be embedded directly into the bearing metal so that a temperature rise will be sensed almost immediately. Each bearing’s temperature is often indicated on a recording instrument, which also signals the alarm and shutdown systems when a temperature rises above safe operating limits.

Vibration detectors come in pairs and are mounted 90 from each other so that vibration in two directions can be seen. Excessive vibration can result from:

Internal failure in the compressor.

Surging.

Misalignment of the compressor and driver.

Another device used for mechanical protection is an axial thrust probe, which measures movement of the shaft along its axis.

The thrust bearing holds the shaft so that it has practically no movement along its axis. However, high thrust forces or worn thrust bearings may result in the shaft moving slightly. The movement is sensed by the probe and an alarm or shutdown would occur.

The compressor can also be protected by radial vibration probes, which measure any radial movement of the compressor shaft.

An audible alarm warns the Control Room Operator if there is unusual shaft movement and in the event of a high displacement indication, a compressor trip is automatically initiated.


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8.0 AXIAL COMPRESSOR DESCRIPTION

The compressor is a multi-stage unit because the pressure increase over each stage is small (one stage is a single set of rotor blades followed by one set of stator blades).

The stages usually multiply the preceding stage's pressure by a value of 110% to 120%. The pressure rise is very small during the first few stages but increases rapidly through subsequent stages because of the multiplying effect. A modern axial-flow compressor can increase the pressure 24-fold in 15 stages.

Figure 8-1

The hub carrying the rotor blades is called the rotor. A series of static blades fixed inside the casing are designed to deflect the gas when it leaves the rotor assembly. This ring of static blades or stator blades acts as a diffuser to decrease the gas velocity. This action turns the kinetic energy into pressure.


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Figure 8-2

From the front to the rear of the compressor (the front being the low pressure end, the rear being the high pressure end) there is a gradual narrowing of the annular space between the rotor and the casing. This is necessary to maintain the axial velocity of the gas as its density increases along the length of the compressor.

Figure 8-3


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8.1 Construction Materials

Axial compressor design was improved with the development of aerodynamics and manufacturing techniques. Blades have complex shapes that were difficult and costly to manufacture before new techniques were developed. New blade materials and designs helped to improve power, cost and weight in the engine as more work could be extracted from the hot air (in the turbine applications) which would increase the power given to the load compressors.

Materials are chosen for their fatigue resistance properties at the temperatures, pressures required and the corrosive nature of the gas used. Weight and strength must also be taken into account. The rotor and stator blades of an axial compressor are usually manufactured from aluminium alloy, steel or titanium alloys, depending upon the location and the maximum temperature. Compressor casings are manufactured from mild steel, stainless steel, magnesium or aluminium alloys and are sometimes lined with a noise absorbent material.

The balancing of the rotor assembly during its manufacture is an extremely important operation. In view of its high operating speeds and the weight of the materials, any imbalance will affect the rotating assembly bearings and compressor operation. To counteract any imbalance, corrections are applied in two planes, radial and axial, usually at each end of the assembly.

A modern multi-stage axial compressor is more efficient, produces a higher pressure and is lighter than a centrifugal compressor. However centrifugal compressors are less expensive, simpler in design and are more robust.


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8.2 OPERATING PRINCIPLES

During operation, the driver turns the rotor at high speed. Gas is continuously pulled into the compressor because of the pressure differential caused by shape of the rotor blades. These rotating blades accelerate the gas which is then swept backwards onto a row of stator blades.

Figure 8-4

The pressure increase results from the diffusion process in the rotor blade passages and a similar process in the stator blades (i.e. kinetic energy is converted into pressure energy). The stator blades also serve to redirect the gas flow at the correct angle onto the next row of rotor blades. The design of the blades in each rotor stage is different due to the change in pressure and volume of the gas.


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Figure 8-5

The last row of stator blades act as ‘straightening vanes’ to remove the gases swirling motion so it will enter the discharge at an even velocity.

A volute type diffuser consisting of an enlarged area is often used downstream of the last stator to further reduce velocity and increase the discharge pressure.

Across each stage, the difference between the inlet pressure and the outlet pressure if very small. The reason it is kept small is that the rate of diffusion and deflection angle of the blades must be limited if losses due to gas breakaway and resultant blade stall are to be avoided.

The small pressure rise through each stage together with the smooth flow path of the gas does much to contribute to the high efficiency of the axial compressor by keeping the gas flow at subsonic speeds.

Compressor blades are designed to produce a given pressure rise and velocity increase over the engine speed range. If something should disturb the pressure/velocity/rotational speed relationship, then the gas flow across the blade profile will break away and create eddies until the blade stalls.

Figure 8-6

Stall is easily understood if you imagine a fan sucking air into a sealed room. Eventually the pressure in the sealed room equals the duty of the fan, although the fan continues to operate it cannot move any more air. The fan has reached its stall pressure. This condition tends to occur on start-up and shutdown when the rotational speed of the compressor is lower than its operational speed.

If the stall condition of a stage or a number of stages continues until all the stages are stalled then the compressor will surge.

Surge is a phenomenon that affects all dynamic compressors. It is a spasmodic or pulsating reversal of flow back through the compressor, basically caused by the


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flow through the compressor being below the minimum stable level.

Surge occurs because one or a combination of conditions exceeds the capabilities of the compressor. These can include:

Power availability below requirements,

Design ratios exceeded,

Flow restrictions.

The change of state from stall to surge can be so rapid as to pass undetected. On the other hand, a stall may be so weak as to produce only slight vibration or poor acceleration. At low speeds or off-design speeds, a slight degree of compressor stalling almost always occurs in the front stages of the compressor, even when a gas flow control system is used.

This condition is normally unnoticeable and will not harm the compressor operation. A severe compressor stall is indicated by a rise in gas temperature, vibration, or a coughing sound. A surge produces high vibration, temperature and a banging sound of varying loudness.

The amount of gas flow and the pressure ratio at which this occurs is called the ‘surge point’. This point is a characteristic of each compressor speed. A line which joins the surge points at various compressor speeds is called the surge line or curve, which is similar to the curves shown for the centrifugal compressors. The surge line defines the minimum stable flow that can be obtained at any rotational speed.

On start-up, it is necessary to use a gas flow control system to prevent surge. The system normally consists of bleeding gas from an intermediate stage to atmosphere or back to the intake. Also, a row or a number of rows of intake guide vanes can be arranged so that their angle can be automatically adjusted to minimise the stalling of the front stage blades at off-design compressor conditions.

Intake guide vanes guide the gas onto the first set of rotor blades. The guide vanes give a swirling motion to the gas entering the compressor suction, they are normally adjustable to control gas flow requirements at various operating conditions.

The amount of gas bled from the compressor is controlled by a hydraulically or pneumatically operated valve, which is open at low speeds and closes automatically when the compressor is at or above its minimum stable flow. When the valve is open, the gas flow across the rear stages is decreased. This prevents the rear stages from ‘choking’ due to the high axial velocity of the gas and the front stages from stalling due to the low flow.

8.3 Associated Problems

Axial compressors rotate at very high speeds e.g. the Solar Saturn gas turbine operates at up to 22,000 r.p.m. and the blades are of a light construction. Any particles entering the compressor can crack, bend, twist or chip a blade resulting in loss of efficiency, high vibration and in extreme cases, destruction of the blades.

Dirt causes gradual erosion, the blade is slowly worn down destroying its aerofoil shape and reducing its efficiency. Dirt can also build up and coat the blade, altering its designed shape, reducing efficiency and and causing an imbalance.


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9.0 LUBE AND SEAL OIL SYSTEM

The lube and seal oil systems for axial flow compressors handling hydrocarbon gas may be separate, or they may be combined into a single system, depending upon the size, type of machine and the type of seal.

Axial flow compressors on gas turbines do not require a liquid seal system as the medium being compressed is atmospheric air. Pressurised seal air from the axial flow compressor is supplied to turbine seals.

Lubrication is required for the shaft journal bearings, thrust bearings, -couplings and gear boxes.

The type of seals used will determine whether a separate seal oil system is required or if the lube oil system can also supply oil to the seals.

9.1 Lubricant

Proper lubrication is the most important compressor operational requirement and only high quality lube oils should be used. Lubricants perform the following functions:

Separate rubbing parts.

Remove heat of friction.

Remove wear material, dirt, and debris from the surfaces.

Provide sealing action.

Protect metals from corrosion.

Most lubricants are refined from a petroleum base. To improve their lubricating properties, chemicals called additives are included that help to improve viscosity and prevent rust/corrosion.

Lube oils do not last forever. Even with additives, lube oil may become corrosive or lose its viscosity because of a build-up of contaminants. Lube oil should be tested regularly to determine its quality and changed when necessary.


Centrifugal & Axialcompressors

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