Screw Compressors

Chapter 6

Screw Compressors Eugene “Buddy” Broerman*, Tim Manthey†, J€ urgen Wennemar‡ and Justin Hollingsworth* *

Southwest Research Institute, San Antonio, TX, United States, †Aerzen USA, Coatesville, PA, United States, ‡MAN Energy Solutions SE, Oberhausen, Germany

Two Types of Screw Compressors Screw compressors generally come in two basic designs: dry screw compressors or wet/oil-flooded screw compressors. Both compressor types consist of rotors that are closely mated. Dry screw compressors do not have any oil in-between the screws. A dry screw compressor consists of two rotors, a closely mating pair that are installed in a tight clearance cylindrical bore. Wet or oil-flooded screw compressors have oil in-between the screws. An oil-flooded screw compressor also has a pair of closely mating rotors that are installed in a tight clearance cylinder bore; however, oil-flooded screw compressors do not require clearances to be as tight as those that are required for the dry screw compressors. The subsequent sections of this chapter describe the various aspects (similarities and differences) of these two types of screw compressors.

Working Principle of Screw Compressors Screw compressors consist of two rotors in a common casing. Both rotors carry intermeshing helical lobes and rotate against each other with tight clearances between the rotors, and between the rotors and casing. During rotation, the lobes and casing form compression chambers that steadily decrease in volume as the rotors turn, changing cyclically from maximum volume to zero and back to maximum again. Thus, the working principle is similar to other positive displacement machines like reciprocating compressors. In an oil-flooded screw compressor, a slide valve is available for capacity control. This slide valve moves axially beneath the rotors and changes the effective rotor length, and also opens an internal recycle volume on the suction side of the compressor. This is explained in more detail in the sections further. Fig. 6.1 shows the rotors in different positions and with a varying working chamber volume. Compression Machinery for Oil and Gas. https://doi.org/10.1016/B978-0-12-814683-5.00006-7 © 2019 Elsevier Inc. All rights reserved.

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FIG. 6.1 Compression progression.

Positive displacement machines do not transfer a certain speed-related amount of kinetic energy to the gas which is converted into pressure like centrifugal compressors but compress the gas “quasi-statically.” The enthalpy difference per mass unit of gas transferred into the gas can vary heavily for different gases. Therefore, the concept of head does not make sense for positive displacement machines and the use of an h-s diagram is not helpful. The working process consists of three phases and is best described in a pressure-volume diagram (Fig. 6.2). (Note that all pressures are absolute pressure unless specifically noted.) Suction phase: The working chamber is connected to the suction line via the inlet port in the casing. The size of the working chamber increases from zero to its maximum value and the chamber is filled with gas at the suction pressure, p1. Compression phase: The working chamber is closed off from the suction and discharge line. The working chamber decreases its size from the maximum value V1 to a defined value V2 that is defined by the following equation: V2 ¼ V1 =vi

(6.1)

with V1 and V2 as defined in the nomenclature section. vi is the built-in volume ratio.

FIG. 6.2 Best fit.

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The built-in volume ratio vi is an important value for the design of screw compressors and is determined by the shape and size of the discharge opening. In oil-free screw compressors, the vi is a fixed value. In oil-flooded screw compressors, some models are available with an adjustable vi slide that is independent of the capacity slide valve; however, even for those without this adjustable vi feature, the vi is only truly realized when the capacity slide valve is in the maximum capacity position. When the slide valve moves toward the minimum capacity position, the effective vi is reduced due to the shorter effective rotor length. Due to the size reduction of the working chamber, the gas is compressed as it is moved by the screws, and the gas pressure and temperature increase. At the end of the compression phase the internal pressure p2i is reached. As described in the following equation, p2i is a function of p1, vi, and k. p2i ¼ p1  vi k

(6.2)

with k being the isentropic exponent of the gas. For ideal gases: k ¼ cp =cv

(6.3)

The ratio p2i/p1, also expressed as vik, is often called the “compression ratio.” It should be noted that the “built-in volume ratio” is a design property of a certain screw compressor, the “compression ratio” depends on the volume ratio and the kappa value of the gas, and the “pressure ratio” (p2/p1) also depends on the discharge line pressure. Therefore, when discussing the pressure ratio for screw compressors, it is important to distinguish whether the discussion is about the “pressure ratio” or the “compression ratio.” Discharge phase: When the compression chamber has reduced its size to the volume V2, the rotor lobes pass the outlet port in the casing and the chamber is connected to the discharge line. This is the beginning of the discharge phase. By further rotation of the rotors the working chamber reduces its size to zero and the gas is ejected into the discharge line at the discharge pressure p2. The screw compressor has no dead volume like a reciprocating compressor therefore no reexpansion of trapped gas happens. During all phases, a small amount of gas leaks across the rotor clearances from the discharge line into the compression chamber. Another leakage flow occurs during compression from the closed compression chamber to the trailing chambers with lower pressure and to the suction line. The leakage to the suction line acts like an internal bypass and reduces the inlet volume flow. This is known as slippage, which also has thermal effects on the compressor: the gas that “slips” backwards from higher pressure to lower pressure has already been heated by the compression process. As it is recompressed, the heat increases further. Thus, higher slippage leads to higher discharge temperature. Slippage increases as internal clearances increase, as pressure ratio increases,

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and as gas molecular weight decreases. All compressor sizing and performance calculations take slippage into account; therefore, it is critical to have accurate gas data and clearance information in order to get accurate data with respect to volumetric efficiency and discharge temperature. Please note that the working process description given above is idealized because the real compression phase is not isentropic but polytropic. The work W needed for the compression is indicated by the area limited by the suction, compression, and discharge line and is mathematically given by. ð W ¼ Vdp (4) It must be noted that the beginning of the discharge is only determined by the position of the outlet ports in the casing and is independent from the actual pressure in the discharge line. Therefore, the internal pressure p2i may deviate from the discharge line pressure p2. This may occur if the suction or discharge line pressures change or if the isentropic exponent k of the gas changes.

Undercompression A case where the internal pressure p2i is lower than the discharge line pressure p2 is called undercompression (Fig. 6.3). In this case the compression chamber opens before the inner pressure p2i has reached the line pressure p2 and gas is flowing back rapidly from the discharge line into the compression chamber until the pressures have equalized. By further rotation of the rotors the volume is finally reduced to zero and the gas is expelled into the discharge line at the pressure p2. Fig. 6.3 shows that the area between the lines is larger than the ideal process in Fig. 6.2 thus indicating that for undercompression a larger compression work W is needed. This is equivalent to a lower efficiency and higher

FIG. 6.3 Undercompression.

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discharge temperature. In addition, the backflow of gas into the compression chamber causes gas pulsations and noise in the chamber and the connected discharge line.

Overcompression The opposite case where the internal pressure p2i is higher than the discharge line pressure p2 is called overcompression (Fig. 6.4). Here the gas is compressed inside the closed compression chamber to a higher pressure than in the discharge line. When the rotor lobes have reached the outlet port edges the chamber opens and the gas expands rapidly into the discharge line until the pressures equalize at p2. By further rotation the chamber volume is reduced to zero and the gas is expelled into the discharge line at the pressure p2. The triangular pressure peak in Fig. 6.4 shows that overcompression also leads to a larger work than in Fig. 6.2 which indicates a higher power consumption and worse efficiency. Again the rapid equalization of gas pressures may cause gas pulsations and noise. The compression to a high internal pressure p2i also leads to high internal gas temperatures which are higher than the temperature in the discharge line. In extreme cases, this may cause damages due to internal overheating that may not be detected by the temperature sensors in the discharge line. In some cases, overcompression can lead to internal pressures that exceed the pressure rating of the machine, and can also cause very high loads on the radial and thrust bearings, leading to reduced bearing life and possibly even shaft damage. It should be noted that Figs. 6.2–6.4 are idealized cases. In reality, the compression phase is not isentropic and also the pressure equalization in case of undercompression or overcompression is not instantaneous but takes some time. Therefore, Fig. 6.5 shows an example of a more realistic simulation.

FIG. 6.4 Overcompression.

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FIG. 6.5 p–v-Diagram simulated.

A certain mismatch of “built-in pressure ratio” and external pressure ratio is often unavoidable and does not cause a real disadvantage. Large deviations between both values, however, may have severe disadvantages such as a drop of efficiency, gas pulsations, and internal overheating in case of overcompression. Undercompression is sometimes unavoidable, especially in oil-flooded screw compressors where high-pressure ratios are possible; cases with an internal compression ratio of 10 and a pressure ratio of 25 have been known to operate without any issues. In any case a moderate undercompression is better than overcompression. The working phases are repeated for each lobe of the male rotor. Thus a compressor with four lobes on the male rotor performs four compression cycles during each rotation of the male rotor. The number of compression cycles per second is called pocket passing frequency (PPF). Gas is discharged into the discharge line discontinuously at the PPF. Table 6.1 shows examples of different compressor types and speed with their respective pocket passing frequencies. The PPF is much higher than the speed of reciprocating compressors which means that the discontinuous flow pulsations occur at higher frequencies. The discontinuous flow requires a careful discharge silencer design.

Comparison of Positive Displacement Machines (Screw Compressor, Reciprocating Compressor) Versus Centrifugal Compressors Screw compressors are positive displacement machines with purely rotary motion.

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TABLE 6.1 Pocket Passing Frequency of Different Screw Compressor Types Small Dry Screw Compressor

Large Dry Screw Compressor

Oil-Flooded Screw Compressor

Male-rotor speed (rpm)

18,000

3000

3600

Number of lobes on male rotor

4

4

5

Number of working cycles per minute

72,000

12,000

18,000

Pocket passing frequency (Hz)

1200

200

300

Compressor Type

The thermodynamic behavior of screw compressors is similar to reciprocating compressors. The power consumption is nearly independent of the gas molecular weight. The inlet gas volume flow does not change much when the pressure ratio or the molecular weight of the gas changes. Therefore, screw compressors do not have a surge line. Due to the insensitivity against mole weight changes a screw compressor can operate with a variety of gases. A screw compressor can operate at very high-pressure ratios as long as the allowable discharge temperature or mechanical limits like bearing loads or shaft stress limits are not exceeded. With oil-flooded screw compressors pressure ratios of as high as 25 in one stage can be achieved. With dry screw compressors a pressure ratio up to 10 is possible with liquid injection. The compact rotor design with a small number of lobes gives a robust rotor. The first lateral critical speed is always higher than the maximum allowable speed (“stiff rotor design”). Due to the purely rotating motion the vibrations of screw compressors are similar to centrifugal compressors and much lower than for reciprocating compressors. While vibration in the machine itself is possible (due to imbalance, bearing failure, torsional resonance, etc.), vibration in a screw compressor package (piping, vessels, baseframe, etc.) is much more likely to come from the effects of gas pulsations than from the motion of the compressor itself. The clearances between the rotors and the casing are only fractions of a millimeter which means that any buildup of process gas residues is scraped away and the balance status of the rotors is not affected. In contrast to reciprocating compressors screw compressors do not have valves for controlling the gas inlet and outlet of the working chamber. The gas inlet and outlet are controlled by ports in the casing. Therefore, failure of valves is not an issue with screw compressors.

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Screw compressors do not have any contacting parts other than shaft seals. Oil-injected screw compressors are typically direct driven by electric drivers or gas engines. Therefore, the typical speeds are between 1500 and 3600 rpm. The speeds of dry screw compressors can range between approximately 1000 rpm and up to 25,000 rpm depending on the machine size. The tip speeds for both screw compressor types are lower than those of centrifugal or axial compressors. Therefore, screw compressors are less susceptible to erosion by droplets or contaminants. With oil-flooded screw compressors, erosion is not an issue due to the relatively low rotor tip speeds. Oil-free screw compressors with continuous liquid injection at the suction side can, due to higher rotor tip speeds, be subject to erosion of the rotors and—to a lesser extent—the casing, if carbon steel is used. Stainless steel rotors are often used in this case, and a stainless steel casing is sometimes used as well.

Differences Between Dry Screws and Oil-Flooded Screws Dry Screw Compressors Fig. 6.6 shows sectional views of a dry screw compressor. A synchronizing gear (often called timing gear) is used to avoid contact between the rotors. In oil and gas service the bearings are typically hydrodynamic journal and thrust bearings. At each shaft end a shaft seal is placed between the compression chamber and the journal bearing (sometimes called conveying chamber seals). Depending on the type of coupling and coupling guard a labyrinth seal for sealing the driveshaft may be necessary. An important value for the characterization of screw compressors is the male-rotor tip speed uM. For dry screw compressors the typical tip speed range for all compressor sizes is between 50 and 150 m/s. In some cases, even lower or higher tip speeds are realized. The rotor diameters are in the range of approximately 100 mm up to >800 mm. Therefore, the largest types have speeds of 1000 to 3600 rpm while the smallest sizes may have speeds up to 25,000 rpm with medium-size compressors in between. Due to the high speeds a gearbox between driver and compressor is needed for small- to medium-sized machines. For large machines a direct drive with a steam turbine is often used but an electric drive with or without gearbox is also common. The built-in volume ratio vi is defined by the casing geometry and is a fixed value for a dry screw compressor. Dry screw compressors perform well with many gases (e.g., corrosive, toxic, flammable), and with changing molecular weights. If required a liquid injection for cooling or washing purposes is possible.

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A

4

1

3

B

5

7

6

5

2

4

8

Key A B

Inlet Outlet

1 2 3 4

Casing Male rotor Female rotor Shaft seal

5 6 7 8

Radial/thrust bearing Timing gear End cover Drive shaft

FIG. 6.6 Axial split dry screw. (From MAN Energy Solutions.)

Oil-Flooded Screw Compressors Figs. 6.7 and 6.8 show sectional views of an oil-flooded screw compressor. In oilflooded screw compressors, sometimes referred to as oil-injected or wet screw compressors, oil is injected directly into the rotor chamber continuously during operation. The oil is discharged with the gas into an oil separator vessel and then must be separated from the gas on the discharge side before being injected back into the compressor again. The injection oil serves several purposes:

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1

A

8

11

9 12

10 B 4

4

2

5

5

7

14

6 3

13

5

5

Key A B

Inlet Discharge

1 2 3 4 5 6 7

Casing Male (main) rotor Female (secondary) rotor Radial bearings Axial (thrust) bearings Oil pump (shaft-driven, optional) Shaft seal

8 9 10 11 12 13 14

Capacity slide valve piston Capacity slide valve vi slide (for variable vi, optional) Capacity slide valve position sensor vi adjustment screw (for variable vi, optional) Oil injection port Thrust balancing piston

FIG. 6.7 Oil-flooded screw main components.

1. It provides a lubricating film between the male and female rotors. Oilflooded screw compressors do not have timing gears. Instead, one rotor drives the other through direct contact (with an oil film between the two rotors). The drive rotor refers to the rotor that is coupled to the motor, while the driven rotor refers to the rotor that is moved by the drive rotor.

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1

263

6

6

5 11

4 9

10

8 7

4

B

3

2

Key A B

Inlet Discharge (not visible)

1 2 3 4 5 6

Casing Male (main) rotor Female (secondary) rotor Radial bearings Axial (thrust) bearings Shaft seal

7 8 9 10 11

Capacity slide valve hydraulic cylinder Capacity slide valve (not visible – under rotors) Capacity slide valve position sensor Oil injection port Thrust balancing piston

FIG. 6.8 Oil-flooded screw main components—3D.

Most oil-flooded screw compressors are male-rotor drive, but many femalerotor drive machines are available as well. 2. It carries away much of the heat of compression. This enables higher pressure ratios than are thermally possible in oil-free screw compressors (in which the compression power increases the gas temperature, and a small part of the heat of compression is absorbed by the casing and rotors). 3. It fills the internal clearances, increasing volumetric efficiency. The volumetric efficiency of an oil-flooded screw compressor can be 10%–20% higher than a similar oil-free screw compressor. 4. It continuously flushes away contamination that might enter the machine from the suction header. There are no internal seals at the conveying chamber of an oil-flooded screw compressor—all of the internal components are in contact with the oil and the gas. Therefore, all of the compressor internals must be compatible with the oil and the gas, and also must be rated for the full range of operating pressures and temperatures. The only seal is at the driveshaft, typically an oilpurged mechanical seal, either single or double. When a single mechanical

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seal is used, the seal is usually fed by the common oil system. When higher safety requirements call for a double mechanical seal, it is typical for a separate oil system to be used for the driveshaft seal. This allows better control of the temperature, pressure, and flow of the oil to the seal, and also ensures that any oil leakage on the atmospheric side of the seal does not contain any process gas. Capacity control is possible via an integral slide valve (refer to Fig. 6.9). The slide valve is located below the rotors and moves axially, actuated hydraulically by the oil system and a system of control valves. Since the slide valve is not truly a valve, it is more accurate and clear to use the terms “load” and “unload” rather than “open” and “close” when referring to its movement. Moving the slide valve toward the unloaded position opens a bypass area on the suction side of the machine, which reduces the volume flow of the machine. Moving the slide valve toward the loaded position closes this internal bypass area and increases the volume flow of the machine. With a single-acting slide valve configuration, the slide valve is moved toward the unloaded position via oil pressure, and is moved toward the loaded position via discharge pressure. Thus, a small pressure difference must exist between suction and discharge in order to move the slide valve toward the loaded position. With a double-acting slide valve, the movement toward loaded and unloaded positions is done by the oil system and is not dependent on discharge pressure.

3

2

1 5

4

Key 1 2 3 4 5

Capacity slide valve Capacity slide valve piston vi cut in capacity slide (shape/length of cut determines maximum vi) Internal recycle (when slide valve is away from maximum capacity position) vi slide (changes location of vi cut in capacity slide relative to discharge port)

FIG. 6.9 Slide valve function.

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Design Features In large screw compressors the casing is normally split horizontally (axially) for ease of maintenance (see Fig. 6.6). For oil-injected screw compressors or for small- and medium-sized dry compressors or for higher discharge pressures, vertically (radially) split casings are common (Fig. 6.10). The suction nozzle is normally on top of the casing while the discharge nozzle is in most cases directed downwards, in order to ensure a free draining of any liquids into the discharge line. Thus the gas flow in the casing is directed from top to bottom and also axially from suction to discharge end. Some screw compressors have top or side discharge arrangements, which offers some advantages in terms of piping layout, but this is not common in process gas machines. The pressure difference between discharge and suction causes not only axial gas forces but also large radial gas forces on the rotors. The radial gas forces cause bending stresses in the shafts and loads on the journal bearings. Journal and thrust bearing forces increase linearly with the pressure difference. The direction of the radial forces changes for different operating conditions. Normally the radial forces are much higher than the rotor weight and therefore robust journal bearings are required. The allowable bearing loads and the shaft stress give a limit for the allowable pressure difference. For example, the thermodynamic process gives the same discharge temperature and efficiency when operating with a pressure ratio 3.0 like compressing from 1 bar abs to 3 bar abs, or from 2 bar abs to 6 bar abs or from 10 bar abs to 30 bar abs. Due to the large pressure difference this machine will be mechanically overloaded when compressing from 10 abs to 30 bar abs.

FIG. 6.10 Vertical split dry screw.

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Rotor Design The rotor design is defined by the shape and number of lobes on the rotors and the rotor length indicated by the L/D ratio. For a given rotor diameter the geometrical volume flow increases proportional with the rotor length. It should be noted that the achievable pressure ratio is independent of the L/D; however, the allowable pressure difference is inversely proportional to L/D. The rotor design is carefully chosen with respect to the intended applications. Any design is a compromise between the conflicting requirements of a large volume flow for a given machine size and high rotor bending stiffness. Compressors for high flow and low-pressure difference have long rotors with few lobes. These are normally dry screw compressors. Lobe combinations may be three lobes on the male rotor and four or five lobes on the female rotor (3/4 or 3/5). L/D is >1.8 up to approximately 2.2. Due to the long and slender rotor these machines are only suitable for low-pressure differences. The most common profile is the asymmetric 4/6 profile (Fig. 6.11). This profile is a good compromise for the conflicting requirements of pressure difference and volume flow for a given size and is used for oil-flooded and for dry screw compressors. Depending on the intended pressure difference the L/D ratio may range from <1 up to approximately 2.2. Fig. 6.12 shows compressors with different L/D ratios. Compressors for high-pressure differences have short stiff rotors with a higher number of lobes. Typical for oil-flooded compressors are lobe combinations 4/6, 5/6, or 5/7 with L/D ratios ranging from approximately 1.2 to 2.8.

Male rotor

Root diameter FIG. 6.11 Rotor profile.

Female rotor

Shaft diameter

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FIG. 6.12 Different L–D.

Dry screw compressors for the oil and gas industry normally have hydrodynamic journal and thrust bearings. The radial bearings are normally sleeve (journal) bearings due to their high load capacity and insensitivity to changes of load direction. Oil-flooded screw compressors for the oil and gas industry also use sleeve radial bearings, and either antifriction or hydrodynamic thrust bearings. For dry screw compressors a shaft seal is placed at each shaft end, between the compression chamber and the bearing. For oil-flooded screw compressors, there are no internal seals, and the only seal is at the driveshaft. This is described in more detail in the oil-flooded screw compressors section above. Different types of seals are available. See separate chapter on seals. At the suction end of a dry screw compressor a synchronizing gear (often called timing gear) with the same gear ratio as the lobe number ratio but normal gear teeth is located.

Shaft Ends and Coupling For direct-driven compressors flexible element couplings are used between driver and compressor. For gear-driven compressors flexible element couplings or torsion shaft couplings (quill shaft) are used to transmit the torque from the gearbox to the compressor (Fig. 6.13). If a torsion shaft coupling is used, the gearbox pinion does not have thrust bearings and the torsion shaft transmits the axial forces from the helical gear and reduces the load on the compressor thrust bearing. Also, the torsion shaft has a low torsional stiffness which means that the first torsional natural frequency of the train is lower than any of the train speeds. Therefore, torsional excitations from the driver cannot be transmitted to the compressor and do not cause torsional stress peaks.

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FIG. 6.13 Dry screw with torsion shaft.

Two-Stage Casing Arrangement Dry screw compressor trains may consist of one single stage or more stages. Normally each stage has its own separate casing. The casings can be arranged in tandem with one stage driving the other directly or in parallel where each stage is connected to a separate pinion of the gearbox (Fig. 6.14). The parallel

FIG. 6.14 Two-stage arrangement for dry screw.

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arrangement has the advantage that the speed of both stages can be chosen to their optimum and thus optimum efficiency is ensured. Also both stages can use their own torsion shaft and have their thrust bearing loads reduced. The tandem arrangement has a simpler gearbox with only one pinion. On the other hand, the speed of the second stage and the interstage pressure cannot be chosen freely and may not result in optimum efficiency. As a further disadvantage the first stage has to transmit the torque of the second stage which results in higher torsional stresses in the driveshaft of the first stage. Therefore, the parallel arrangement is a more versatile design.

Typical Shaft Seal Arrangements Shaft Seals for Dry Screw Compressors The purpose of shaft seals is to prevent leakage of process gas to the bearings and the oil system and to the atmosphere. In vacuum compressors shaft seals must prevent air from entering the compression chamber. The optimum seal choice depends on the actual process conditions (dry or liquid injected, seal pressures, availability of suitable seal medium, compatibility of seal medium with the process) and of course economic considerations. The seal type should be chosen by the compressor manufacturer together with the process engineer. Typically, dry screw compressor seals are one of these types: l l l l l l

Carbon ring seals in various arrangements Liquid injected restrictive ring seals Oil-cooled mechanical seals Double or tandem dry gas seals Labyrinth at driving shaft A combined seal incorporating multiple of the above elements

Fig. 6.15 shows a simple carbon ring seal arrangement where the discharge end seal is balanced to the suction line. The suction end seal and the outboard end seal on the discharge end are buffered with nitrogen which prevents process gas from escaping into the oil reservoir. This arrangement is very simple and cost effective because it needs only one nitrogen control valve with pressure monitoring and nitrogen pressure and flow are relatively low. If the nitrogen supply fails gas will escape into the oil reservoir. Fig. 6.16 shows a more complex carbon ring seal with two different seal gas supplies. Buffer gas is injected at suction and discharge end with pressure control above suction and discharge pressure, respectively. One part of the buffer gas enters the process chamber and the other part is routed to a flare. The buffer gas may be nitrogen but other gases like natural gas or steam could also be used. The outboard seal acts as a separation seal and is fed with nitrogen controlled above flare pressure. Carbon ring seals are simple, cost effective, and reliable seals. They are typically used in applications with pressures <1 MPa. Nitrogen is required as

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FIG. 6.15 Balanced carbon ring seal.

FIG. 6.16 Carbon ring seal with separation.

sealing medium. If the buffer gas fails, the compressor must be depressurized immediately in order to prevent leakage of process gas. The carbon rings are not gastight and therefore a certain leakage to flare or to the oil reservoir is unavoidable during depressurization. Fig. 6.17 shows a water buffered restrictive ring-type seal. This is a good choice for dry screw compressors operating with dirty and dusty gases like coke oven gas, lime kiln gas, or acetylene crack gas which require injection of large amounts of demineralized water for cooling and washing purposes. In these cases, the discharge temperature is somewhat below saturation temperature so that a certain amount of water leaves the compressor in the liquid form.

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FIG. 6.17 Water cooled restrictive ring seal.

Demineralized water is fed at the suction and discharge end seals at pressures above suction and discharge pressure, respectively. A part of the water enters the compression chamber and mixes with the process gas. The other part flows to the drain and may be routed back to the water tank. The separation seal is purged with nitrogen at approximately 20 kPa gage and prevents water from entering the oil system. This is a simple, reliable, and cost effective seal but of course it can only be used in cases where ingress of water into the process is allowable. If the seal water fails, the compressor must be depressurized immediately in order to prevent process gas leakage. The restrictive rings are not gastight and therefore a certain leakage to the flare is unavoidable during depressurization. If the nitrogen fails, the compressor must be shutdown also in order to prevent water leaking into the oil reservoir. For cases where no seal water or sealing gas is available an oil-cooled mechanical seal is a good choice. Fig. 6.18 shows a single oil-cooled mechanical seal and Fig. 6.19 shows a double oil-cooled mechanical seal. In both cases the seal is a combination of carbon ring bushings with mechanical seals. The mechanical seal consists of a stationary ring and a rotating ring which form a narrow seal gap extending in radial direction. Both rings are pressed together by springs and by hydraulic forces and slide against each other while a thin oil film separates the faces. Due to the sliding motion of the seal faces heat is generated. Oil with a pressure higher than gas pressure is injected into the annulus at the bottom and cools the seal faces. Also the oil seals the gap between the seal faces and prevents gas from escaping. The differential pressure between oil and process gas is typically around 0.2–0.3 MPa and is controlled by a differential pressure control valve at the inlet to the annulus. The outflow of oil is

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FIG. 6.18 Single mechanical seal.

FIG. 6.19 Double mechanical seal.

limited by orifices at the outlet. In case of a single mechanical seal the outboard oil flow is limited by a restrictive ring. The main part of the oil leaves the annulus either on top of the chamber or in the outboard direction through dedicated orifices. A small amount of oil leaks in the inboard direction. This leak oil is routed together with gas from the carbon ring bushings to an oil separator, where the oil is separated and the gas is fed

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back into the suction line. Normally the leak oil is polluted by the process gas and therefore cannot be reused. The discharge end seal is balanced to the suction line. Thus both seal ends can be buffered with seal oil at the same pressure level. Normally lube oil and seal oil are the same and have one common oil reservoir. Therefore, the requirements of the seals must be checked when choosing an oil type. The oil must not contain zinc additives because these tend to form layers on the faces of the seals. A combined seal is any seal that incorporates more than one primary seal type/medium into a single seal arrangement. The water buffered restrictive ring seal mentioned above is technically a combined seal because it includes a separation seal that is purged with gas, which prevents sealing water from reaching the oil, and lubrication oil from the bearings from reaching the water return connections of the compressor. Another common type of combined seal is a gaspurged carbon ring seal on the process side with an oil-purged mechanical seal on the oil side. The oil-purged seal reduces the consumption of seal gas from the carbon ring seals and provides positive shut-off during standstill, while the gaspurged carbon ring prevents oil from the mechanical seal from contaminating the process. Mechanical seals are a reliable and cost effective seal type which can be used in a variety of applications. In case of seal oil failure, the compressor must be shutdown immediately. Failure of seal oil supply may damage the seal faces, but if the seal faces remain intact the seal will contain the gas during standstill without gas leakage to the atmosphere. For dry screw compressors mechanical seals are normally used up to 1.5 MPa suction pressure although higher pressures are possible. It should be noted that the oil system will be more expensive due to larger oil flow and higher oil pressure than for other seal types. The power losses of the mechanical seals are similar to those of the bearings which increases power consumption compared to the other seal types. Dry gas seals are typically arranged in double or tandem design. Fig. 6.20 shows a double dry gas seal. The basic arrangement is the same as for the double mechanical seal but the seal medium in this case is nitrogen. Nitrogen is fed into the annulus between the two seal faces at a pressure of 2–3 bar above the gas pressure. A very small amount of nitrogen leaks via the inboard seal face into the process and via the outboard seal face to a safe location. A separation seal consisting of carbon rings is fed with nitrogen at approximately 20 kPa gage and prevents bearing oil from polluting the dry gas seal. The dry gas seals are separated from the process by carbon ring bushings. The discharge end seal is balanced to the suction line. This seal can only be used, if nitrogen is available at a pressure higher than the highest possible suction pressure [including settle-out pressure (SOP)]. The nitrogen consumption of double dry gas seals at high pressure is much less than for carbon ring seals. Dry gas seals are very sensitive against pollution or

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II Types of Equipment

FIG. 6.20 Double DGS.

liquids, therefore this seal arrangement is not suitable for dirty or wet gases. The seal nitrogen must be filtered carefully and must be dry. If no high-pressure nitrogen is available a tandem dry gas seal as per Fig. 6.21 may be used. Here two dry gas seals are placed in a series. A clean and dry buffer gas is fed into the annulus upstream the primary seal and prevents the process gas from entering the dry gas seal. The largest part of the buffer gas

FIG. 6.21 Tandem DGS.

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flows inboard via the carbon rings and mixes with the process gas. A very small amount of buffer gas passes the primary seal in the outboard direction. A part of the leak gas flows to the primary vent. The other part passes the secondary seal and flows to the secondary vent. In some cases there is a requirement for an intermediate labyrinth between primary and secondary seal which is buffered with nitrogen at the inlet of the secondary seal. Due to the complex casing geometry and axial space restrictions this is in many cases not possible for screw compressors. Between the secondary seal and the bearings there is a separation seal purged with nitrogen at 20 kPa gage (same as for the double dry gas seal) which prevents oil from polluting the dry gas seals. Tandem seals are longer in axial direction than double dry gas seals and therefore a balance line to suction may not be possible due to space restrictions. In this case the buffer gas must be supplied at a pressure higher than discharge pressure. An independent supply of filtered and dry buffer gas is necessary. Tandem dry gas seals are the best choice for high seal pressures (>2 MPa) provided that the required buffer gas quality and availability can be maintained at all times. Typical buffer gases are natural gas or light hydrocarbons. Nitrogen is of course the safest choice if available. Both double and tandem dry gas seals are very sensible against pollution. Ingress of dirt or liquids may lead to immediate damage of the seals. The buffer gas must be filtered to 3 μm. The gas temperature should be at least 20 K above the gas dew point at the actual pressure. Both seal types and the seal support systems are very expensive compared to other seal systems. It must be noted that any kind of seal can work properly only if the sealing medium is supplied at the required pressure, temperature, and quality. Failure of the seal medium supply or failure of the seal causes a drop in supply pressure which is monitored with alarm and trip. The compressor must be tripped by low sealing medium pressure and depressurized to atmospheric pressure to prevent process leakage. If in case of a seal failure the compressor is not depressurized this may lead to a process gas leakage through the seals and into the oil reservoir. The oil reservoir is vented to a safe location. The sizing of the vent line must consider the seal failure case and must be sufficient to maintain the pressure inside the oil reservoir at an acceptable level. Any seal with nitrogen flow to the bearing drain has the positive side effect that the oil reservoir is automatically inertized and can therefore even in case of a seal failure not contain an explosive gas mixture. The shaft seal at the driving shaft of screw compressors separates the bearing compartment from the atmosphere. Inside this compartment there is air with oil mist at atmospheric pressure. Therefore, a simple labyrinth is sufficient. In some cases, a purge with instrument air or nitrogen at a pressure slightly above atmosphere may be used to prevent oil migration along the shaft. The driveshaft seal of an oil-free screw compressor is typically either a labyrinth seal or a single mechanical seal, oil purged. Since the internal seals at the conveying chamber described above can be designed with the level of safety

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and reliability desired, the driveshaft seal is rarely regarded as a primary protection to keep process gas from reaching the atmosphere. It would require multiple concurrent failures for the process gas to reach the driveshaft seal, so normally this seal’s function is only to keep atmospheric air out.

Driveshaft Seals for Oil-Flooded Screw Compressors As described above, there are no internal seals at the conveying chamber of an oil-flooded screw compressor. The only seal is at the driveshaft. Depending on the design of the machine, the internal side of the driveshaft seal may be at suction pressure, discharge pressure, or some intermediate pressure. Regardless, the seal must be designed for the maximum allowable working pressure (MAWP) of the machine, and should provide a positive shut-off even without the presence of seal oil. One option is to use a single mechanical seal, purged with oil from the common oil system. Since the seal typically requires oil at the same pressure and temperature as the rest of the oil system, this is a simple arrangement. The oil outlet from the seal is routed internally to the discharge flange of the machine, so there is not a separate oil outlet connection on the machine. At the outlet of the machine, the seal oil mixes with the rest of the oil being injected into the bearings and rotor chamber, along with the discharge gas, and is routed to the discharge bulk oil separator. From there, the oil is pumped to a higher pressure, cooled, filtered, and reinjected into the machine, and the cycle continues. Since the oil always carries a small amount of process gas in it, the seal materials must be compatible with the process gas. The seal must also be designed to minimize internal pressure drop—if the pressure drop is too high, the gas will flash out of the oil, which can damage the seal. As with any mechanical seal, there is always some amount of leakage to the outboard side. In this arrangement, the leakage oil carries a small amount of process gas with it. Normally the amount of gas coming from the leakage oil is so small that it cannot even be detected by a gas monitor, and in many cases this amount of leakage is acceptable to the plant operator. However, in the event of a seal failure, the amount of leakage oil—and thus the amount of gas carried by the leakage oil—increases, and when the oil system is shutdown, the process gas inside the machine may also leak out at a higher rate than what is acceptable. When this risk is not acceptable to the user, then a double mechanical seal is used. When a double mechanical seal is required, it is most common to also require a separate oil system for the seal. By using a separate oil system, it can be ensured that the seal oil is not in contact with the process gas. There is still outboard seal leakage, but that leakage oil does not carry any process gas with it. There is a small amount of inboard leakage as well, at a quantity that is not enough to significantly alter the properties of the primary oil system. However, the seal oil (which may be a different viscosity than the primary lube oil) must still be compatible with the primary lube oil, with the process gas, and

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with all of the internal components of the machine. With this arrangement, there is a separate oil outlet connection on the machine, through which most of the injected seal oil returns to the reservoir of the separate seal oil system, where it is pumped to a higher pressure, cooled, filtered, and reinjected into the machine, and the cycle continues. By observing the oil level in the reservoir, the total oil leakage can be monitored. The outboard leakage can be directly measured via a separate container. The inboard leakage is simply the difference between the total oil leakage and the measured outboard leakage. Thus, the health of both seals can be monitored independently. When the leakage of either the inboard or outboard seal exceeds the allowable limit, it will be indicated by the loss of oil in the reservoir, which should trigger a shutdown of the machine. The double mechanical seal must be designed such that either of the two seals can individually provide a positive shut-off when the other fails.

Thermodynamic Behavior Suction Flow and Power Consumption Versus Compressor Speed Figs. 6.22 and 6.24 are calculated for two dry screw compressors of 50% and 100% rotor diameter compressing crack gas with constant suction pressure 100 kPa abs and discharge pressure 500 kPa abs with water injection. Fig. 6.22 (and Fig. 6.23) show the suction volume flow versus male-rotor tip speed at constant suction and discharge pressure. The suction volume flow increases linearly with speed and increases with the square of the rotor diameter. Fig. 6.24 (and Fig. 6.25) show the power consumption and the torque vs rotor tip speed for two rotor diameters. In Fig. 6.24 the lines with circles show the power consumption with the scale on the left side and the lines with triangles show the torque with the scale on the right side. The solid lines show the power

Suction volume flow (m3/h)

80,000 70,000

100% rotor diameter

60,000

50% rotor diameter

50,000 40,000 30,000 20,000 10,000 0 40

60

80

100

Male rotor tip speed (m/s)

FIG. 6.22 Volume versus speed for dry screw compressors.

120

140

9000 8000

100% rotor diameter

Suction volume flow (m3/h)

50% rotor diameter

7000 6000 5000 4000 3000 2000 1000 0 32

48

Male rotor tip speed (m/s)

FIG. 6.23 Volume versus speed for oil flooded screw compressors.

FIG. 6.24 Power and torque versus tip speed for dry screw compressors.

FIG. 6.25 Power and torque versus tip speed for oil flooded screw compressors.

65

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and torque for a machine with 100% rotor diameter, and the dotted lines show the power and torque for a machine with 50% rotor diameter. For constant pressures the power consumption increases nearly linearly with speed and the torque is nearly constant. At constant speed the power consumption increases with the square of the rotor diameter and the torque with the cube of the rotor diameter. As a rule of the thumb 50% speed results in 50% suction flow and saves 50% power. Therefore, speed variation is a very good method of process control. The upper limit for speed variation is given by the first lateral critical speed of the rotors including a safety margin. The lower limit is not a fixed value but is normally given by a drop of efficiency and an increase in discharge temperature. With a slide valve available for capacity control in oil-flooded screw compressors, the difference in power consumption between capacity control via slide valve and via VFD should be carefully considered. Capacity control via VFD is almost always more efficient than by using the slide valve (refer to Fig. 6.35). Although the exact difference varies by application and machine, it is always true that the greater the turndown from 100% flow, the greater the disparity in efficiency between VFD and slide valve. This is due to the fact that when using the slide valve for capacity control, the machine is still running at full speed—this uses a certain amount of energy regardless of how much gas is flowing through the machine. When deciding whether to use the slide valve or a VFD for capacity control in a new installation, consideration should be given to how far from 100% flow the machine will be operated, how often, and for how long. TCO, maintenance, and operating simplicity should also be considered.

Actual Suction Volume Flow Versus Pressure Ratio Fig. 6.26 shows the suction volume flow versus discharge pressure. The same characteristics is valid for suction volume flow versus pressure ratio p2/p1. Compressor speed and gas molecular weight are kept constant. The pressure ratio may change by variation of the suction pressure or of the discharge pressure. In both cases the suction volume flow drops slightly with increasing pressure ratio but there is no surging. The limitation in pressure ratio is given by discharge temperature, mechanical limits like bearing load or shaft stress, overcompression or undercompression, and efficiency drop. Fig. 6.27 shows the suction volume flow versus discharge pressure while varying the vi in an oil-flooded screw compressor. The reduction in flow as discharge pressure increases is due to slippage. Although the injected oil helps to fill internal clearances and reduce slippage, it still occurs to some extent, and a lower vi equates to more slippage, although usually by only a few percent. The vi is normally selected based on pressure ratio; however, here it is clear that vi also plays a role in the volume flow.

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Suction volume flow (m3/h)

9000 8000 7000 6000 5000 4000 3000 100% tip speed 75 % tip speed 50% tip speed

2000 1000 0 1

1.5

2

2.5

3

3.5

4

4.5

5

Discharge pressure (bar abs) FIG. 6.26 Flow versus discharge pressure for a dry screw compressor.

FIG. 6.27 Flow versus discharge pressure for an oil flooded screw compressor.

Power Consumption Versus Discharge Pressure Fig. 6.28 shows the power consumption versus discharge pressure with different built-in volume ratios vi. Compressor speed, gas molecular weight, and suction pressure are kept constant. For constant suction pressure the power consumption increases linearly with the discharge pressure. For a small built-in volume ratio, the slope of the power curve is steep resulting in a rapid growth of power consumption with increasing pressure ratio. For a large built-in volume ratio vi, the slope is much less than for a small vi but on the other hand the power

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Power consumption (kW)

400

v i = 1.25

300

v i = 2.0 v i = 2.5 v i = 3.15

200

100

0 1

1.5

2

2.5

3

3.5

4

4.5

5

Discharge pressure (bar abs) FIG. 6.28 Power versus discharge pressure for a dry screw compressor.

FIG. 6.29 Power versus discharge pressure for an oil flooded screw compressor.

consumption at pressure ratio 1 (discharge pressure ¼ suction pressure) is high. Therefore, compressors with large built-in volume ratio are well suited for large pressure ratios but have a high torque in unloaded operation, for example, during start-up. With oil-flooded screw compressors, the power curves with relation to discharge pressure are less steep in the normal operating range. This is because oilflooded screw compressors are generally used for higher pressure applications than oil-free, and there is less inertia-based loss as a percentage of total power consumption (refer to Fig. 6.29).

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Actual Suction Volume Flow and Power Consumption Versus Molecular Weight for Dry Screw Compressors The main thermodynamic property for the compressed gases is the molecular weight. Gases compressed in dry screw compressors range from hydrogen, natural gas, air, hydrocarbons, CO2, and butadiene up to very heavy gases with a molecular weight of 100 kg/kmol. The molecular weight and pressure ratio determine the optimum tip speed. For heavy gases or low-pressure ratio the optimum tip speed is low, for light gases or high-pressure ratio the optimum tip speed is high. Dry screw compressors operating with different gases are compared using the circumferential Mach number at the inlet conditions of the gas. Typical circumferential Mach numbers range between 0.2 and 0.4 although higher or lower values are possible for some applications. Figs. 6.30 and 6.32 show the effect of different molecular weights on the performance of two dry screw compressors. The data have been calculated for a male-rotor tip speed of 120 m/s with a pressure ratio of 3 with water injection. The gases in this example range from a mixture of 50% hydrogen and 50% methane with a molecular weight of 9 kg/kmol up to CO2 with a molecular weight of 44 kg/kmol. Figs. 6.30 and 6.32 show that the suction volume flow and the power consumption do not change very much within a large range of molecular weights. This means that dry screw compressors are suited well for variable gas applications like flare gas compression. For these applications screw compressors are better suited than centrifugal compressors because they do not have surge line. For very low molecular weights the drop in volume flow is more pronounced than the drop in power consumption thus leading to an increase in discharge temperature. If the discharge temperature can be limited by liquid injection even larger variations of molecular weight are possible. The final limitations must always be determined on a case-by-case basis.

Suction volume flow (m3/h)

60,000 50,000 40,000 100% rotor diameter

30,000

50% rotor diameter

20,000 10,000 0 0

10

20

30

Molecular weight (kg/kmol)

FIG. 6.30 Volume versus mole weight for dry screw compressors.

40

50

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The second important gas property is the isentropic exponent k. A high isentropic exponent leads to high discharge temperatures due to internal compression as described in Chapter 2. If the discharge temperature exceeds the design limits of the compressor, then a liquid injection into the suction nozzle may be used to cool the gas.

Actual Suction Volume Flow and Power Consumption Versus Molecular Weight for Oil-Flooded Screw Compressors Figs. 6.31 and 6.33 demonstrate these relationships for oil-flooded screw compressors. The 50% rotor diameter machine in this example has a lower tip speed limit than the 100% rotor diameter machine, so the flowrate is quite a bit lower. 10,000

Suction volume flow (m3/h)

9000 8000 7000 100% rotor diameter 6000

50% rotor diameter

5000 4000 3000 2000 1000 0 10

20

30

40

50

Molecular weight (kg/kmol)

FIG. 6.31 Volume versus mole weight for oil flooded screw compressors.

Power consumption (kW)

3000 2500 2000 100% rotor diameter

1500

50% rotor diameter

1000 500 0 0

10

20 30 Molecular weight (kg/kmol)

FIG. 6.32 Power versus mole weight for dry screw compressors.

40

50

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FIG. 6.33 Power versus mole weight for oil flooded screw compressors.

However, this is a good representation of the consistency of these behaviors. It can be seen that for both machine sizes, the power consumption and flowrate are almost completely independent of the gas mole weight.

Power Consumption Versus vi for Oil-Flooded Screw Compressors As mentioned previously, some oil-flooded screw compressors have a fixed vi, while others have the ability for the vi to be varied. In either case, it is important to consider the working vi of the machine in the context of the working pressure ratio of the machine. The vi should be selected or set such that the final internal pressure of the machine matches the discharge header pressure as closely as possible. Fig. 6.34 demonstrate this for oil-flooded screw compressors. In addition to the impact on power consumption, a large mismatch between internal and external pressures can lead to operating problems as mentioned previously in the Overcompression section.

FIG. 6.34 Power versus vi for oil flooded screws.

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FIG. 6.35 Power, slide valve versus VFD for oil flooded screws.

Operational Guidelines Capacity Control of Dry Screw Compressors The most efficient capacity control method for dry screw compressors is speed control, because the power consumption decreases linearly with the gas volume flow and the speed. A typical turndown value may be between 40% and 60% of rated flow. The lower values are valid for compressors with liquid injection while the higher values are valid for compressors without liquid injection. If speed control is not possible or the required flow is below the minimum allowable speed the capacity is controlled by a cooled recycle line. With a cooled recycle turndown to zero is possible. During recycling no power is saved and therefore this method should be used only in cases where the compressor operates at full capacity for most of the time. Typical input properties for the capacity control are either suction pressure or discharge pressure.

Capacity Control of Flooded Screw Compressors Oil-flooded screw compressors have an internal slide valve as described above. Turndown to 10% or 20% of maximum flow is possible via the slide valve. Operating at reduced capacity with the slide valve means less internal compression, and there is also a certain amount of mechanical and dynamic loss when operating at full speed but reduced slide valve capacity. Thus, using a VFD for capacity control is more efficient than using the slide valve, as discussed in Suction flow and power consumption versus compressor speed above. As with oilfree screw compressors, consideration must be given to the amount of time the machine will be operating at reduced load versus full load.

Starting of Dry Screw Compressors Starting of dry screw compressors should always be unloaded, that is, with a discharge pressure close to suction pressure. The compressor is separated from

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the downstream system by a check valve and a large unloading valve is placed in the recycling line. The unloading valve recycles the full compressor flow from the discharge side to the suction side. During start-up the pressure drop in the recycle line is max. 50 kPa (discharge pressure  suction pressure + 50 kPa) and therefore the starting torque is less than during normal operation. Also the mechanical loads on the bearings during start-up are much lower which prevents metal to metal friction in the bearings at low speeds. After the compressor has reached the desired speed the unloading valve is closed and the discharge pressure begins to rise. Finally, the discharge pressure exceeds the pressure in the downstream system, the check valve in the discharge line opens and the compressor delivers gas into the system. It should be noted that after the unloading valve has been closed gas is sucked out from the suction volume and the suction pressure may drop unless enough gas is delivered from the upstream system. The drop of suction pressure can be reduced by a large suction volume.

Starting of Flooded Screw Compressors Similar to oil-free compressors, oil-flooded compressors with sleeve bearings must start unloaded. In most cases, unloaded start can be accomplished by moving the slide valve to the minimum flow position. Opening the recycle valve can be helpful but is not always necessary. As with any hydrodynamic bearings, prelubrication is required to prevent bearing damage. This is accomplished with a separate oil pump driven normally by an electric motor and sometimes by a steam turbine. The flow of oil required for prelubrication is normally 15%–25% of the fully flow required. One option is to use a small prelube pump and have the main oil pump be shaft driven. Another option is to have a separate 100% oil pump that provides the oil for prelubrication as well as normal operation. To prevent filling the compressor body completely with oil, automated shut-off valves block the flow of oil to the main injection ports in the rotor chamber, and the oil only flows to the bearings and driveshaft seal. Since the pressure requirement for the oil is the same during prelube as during operation (normally 250–320 kPa higher than gas discharge pressure) but the required flow is much lower, the oil system must be designed to handle both prelube and normal operation. Once the compressor has reached minimum speed, the prelube pump (if used) is turned off. The recycle valve can be closed, and the slide valve can move to increase the flow capacity of the machine. The suction and discharge check valves will open and gas is moved downstream.

Stopping of Dry Screw Compressors and Settle-Out Pressure During shutdown dry screw compressors are also unloaded. Immediately with the driver stop signal the unloading valve opens and the gas in the discharge line expands into the suction line. The discharge pressure drops rapidly and the

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check valve at the discharge of the compressor closes. The pressures at compressor discharge and suction equalize after a few seconds at the SOP. The compressor should come to standstill only after the pressures have equalized in order to prevent reverse rotation. Typical run-down times may range from a few seconds for small high-pressure compressors up to approximately 1 min for large compressors operating in vacuum service. The compressor seal system must be able to operate at the SOP and must adjust to the SOP within seconds in order to avoid malfunction of the suction end seals and possible process gas leakage. For compressors with a built-in volume ratio > 2.5 the torque increases with increasing suction pressure. A massive increase of suction pressure may lead to overloading the compressor and coupling. Therefore it is important that the SOP is as close to the suction pressure as practical. The SOP depends on suction pressure, discharge pressure, and the gas volumes at suction side and discharge side up to the check valve. For start-up and shutdown the discharge side gas volume should be as small as practical and the suction end volume should be as large as practical. Check valves in the suction line must be avoided. Block valves in the suction line should be closed only after the pressures have equalized at SOP. If all these measures are not sufficient to reduce the SOP to an acceptable level a quick opening and sufficiently sized blow down valve to a flare should be placed in the discharge line.

Stopping of Flooded Screw Compressors and SOP Oil-flooded screw compressors must also be unloaded during coast down to stop, to protect the bearings. However, it is also important to avoid a rapid depressurization of the discharge side. The oil in the bulk oil separator, any secondary separator vessels, and everywhere in the oil system is at discharge pressure or higher, with some amount of process gas dissolved in the oil. Rapidly reducing the pressure in the system will cause the gas to rapidly bubble out of the oil, which can lead to foaming of the oil (champagne bottle effect). It can take many hours for the foam to return back to liquid form. If undetected, foaming in the oil can damage the compressor if the system is running. The rapid drop in the pressure can also cause oil migration toward the lower pressure area, resulting in a total loss of oil from the bulk oil separator vessel. It is important to have a check valve on the suction side as well as the discharge side, and for the suction check valve to be as close as possible to the compressor. This ensures that the SOP is only slightly lower than the discharge pressure. During coast down, the slide valve should be moving quickly to the unloaded position to protect the bearings from high load at low speeds. If the recycle valve connects downstream of the suction check valve, it can be opened during coast down. Otherwise, it should remain closed. After the compressor has stopped, if it is necessary to reduce the system pressure, it should be vented slowly to prevent oil foaming.

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Liquid Injection for Dry Screw Compressors Liquid injection is essential for oil-injected screw compressors and is often used for dry screw compressors as well. Dry screw compressors use liquid injection for cooling and washing purposes. The most common injection liquid is demineralized water but other liquids like methanol or heavy hydrocarbons may also be used if they are compatible with the process. Fig. 6.36 shows the effect of water injection on the discharge temperature of a large dry screw compressor at two different discharge pressures. Two modes can be identified which are separated by the saturation temperature. The saturation temperature depends on the humidity of the suction gas and the discharge pressure.

Injection Above Saturation Temperature For low injection flows there is a sharp linear decrease of discharge temperature with increasing injection flow. The cooling effect is caused by evaporation of the injected liquid during compression. This characteristic is valid until the saturation temperature is reached. In this mode the injection flow is controlled by a control valve to adjust a defined discharge temperature. Due to temperature measurement precision this mode is practically limited to a discharge temperature of at least 10 K above saturation temperature. This injection mode is used if the discharge temperature must be reduced due to process or compressor design limits.

250

Discharge temperatrure with water injection [°C]

Discharge pressure 5 bar abs

200 Discharge pressure 3 bar abs

150

100

50

0 0

2000

4000

6000

Injection water flow [kg/h] FIG. 6.36 Temperature versus injection flow for a dry screw compressor.

8000

10,000

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Injection Below Saturation Temperature By increasing the water injection flow the discharge temperature eventually reaches saturation temperature and no more water will evaporate. In this mode even a massive increase of injection water flow leads only to minor reduction of discharge temperature. In this case the injection flow cannot be controlled by discharge temperature but is controlled by a flowmeter and a control valve at a predefined flow. Injection below saturation temperature is often used for processes with dirty gases like acetylene crack gas, lime kiln gas, or coke oven gas. A part of the water does not evaporate but leaves the compressor in liquid form and washes the rotors, casing, and piping. For these processes it should be noted that a practical discharge temperature is only approximately 1 K below saturation temperature. In case of polymerizing gases solvents may be injected in addition to water to prevent sticking of the rotors or clogging of piping components.

Liquid Injection Flows for Dry and Oil-Injected Screw Compressors/Liquid Hammer The quantity of injected liquids differs very much between oil injected and dry screw compressors. For dry screw compressors the volume flow of liquid is determined by the evaporation of the liquid and is normally <0.01% by volume of the gas flow at suction conditions. In cases with injection below saturation temperature the injection flow may be 0.01% by volume or higher. For oil-flooded screw compressors the cooling effect depends on the heat transfer between gas and oil. Therefore the percentage of oil compared to gas volume is much larger than in dry screw compressors. Fig. 6.37 shows the relative relationship between gas discharge temperature and oil injection flow. It can be see that in general, the relative cooling effect of the oil decreases

FIG. 6.37 Temperature versus oil injection flow for an oil flooded screw compressor.

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as injection quantity increases. This is an indication that a certain amount of heat will always be taken by the gas and the compressor materials themselves. Oil-flooded screw compressors are designed in such a way that the injected oil leaves the compressor discharge port along with the gas. The amount of oil to be injected depends on the load and the rotating speed. Too much oil injection can cause a “choke” effect at the discharge, increase dynamic losses (reflected by higher shaft power), and increase vibration. Too little oil injection is generally reflected by higher discharge temperature. In some cases, as with gases with water vapor content, it is desirable to keep the discharge temperature as high as possible to stay above the dew point. In these instances, the amount of oil injection must be controlled to maintain the desired discharge temperature. Both compressor types tolerate liquid injection well as long as the liquid comes in a steady and controlled flow. An uncontrolled inrush of a large amount of liquid like an overflow of a separator must be avoided by any means because liquids are not compressible and thus the size reduction of the compression chamber will cause extreme forces and torques on the rotors. This effect is called “liquid hammer” and can cause massive damages of rotors, bearings, seals, and inner casing surfaces. The authors have never experienced a bursting of the steel or ductile iron casing, and this is not expected because the high torque in this case causes a failure of the coupling or the driving shaft and the compressor comes to an immediate standstill.

Recommended Instrumentation and Monitoring For dry and oil-flooded screw compressors the instrumentation listed in Tables 6.2–6.5 is recommended. It should be noted that additional monitoring may be required depending on the application. Shaft vibration monitoring can give much information when interpreted by an experienced vibration specialist. It should be noted however that shaft vibration monitoring is much more expensive than casing vibration monitoring because 8 x–y probes (4 at each rotor) plus 2 key phasors are required. Therefore, x–y probes should be used only when the interpretation of vibration data is required but not if only automatic shutdown is required.

Typical Design Range (Discharge Pressure, Pressure Ratio, Volume Flow, Driver Power, Molecular Weight) Design Range for Dry Screws The maximum discharge pressure for dry screws is limited by the casing design. Typically, the casings are not designed for the individual application but are designed for certain flange pressures (e.g., Class 150, Class 300, 1.6, 2.4, 4 MPa, etc.) Vertical split casings have higher pressure limits than horizontal split casings because it is easier to seal the vertical split including a fully

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TABLE 6.2 Standard Instrumentation for Dry Screw Compressors Trip Low

Suction Pressure

Alarm Low

Suction pressure

Alarm high in order to prevent overcompression or high torque

Trip high

Discharge pressure

Alarm high

Trip high

Suction temperature

Alarm high

Trip high

Discharge temperature

Alarm high

Trip high

Thrust bearing temperature

Alarm high

Trip high

Journal bearing temperature

Alarm high

Trip high

Casing vibration or shaft vibration monitoringa

Alarm high

Trip high

Injection flow (if applicable, for machines that have liquid injection in all specified operating conditions)

Alarm low

Trip low

Lube oil pressure

Alarm low

Trip low

Lube oil pressure

Alarm high

Lube oil temperature

Alarm high Start-up interlock low

a

Due to the stiff casing design, 1 accelerometer per compressor or 1 accelerometer at each compressor end (max 2 in total) is sufficient for automatic monitoring with shutdown.

enclosed O-ring. Dry screw compressors with vertical split and discharge pressures of >5 MPa have been built. Horizontal split casings generally have design pressures up to 1.6 MPa. However, it is important to note that in many cases the compressor may not be able to operate at the maximum discharge pressure of the casing due to other limitations. The maximum discharge pressure must be clearly distinguished from the maximum allowable differential pressure. The maximum allowable differential pressure depends on the rotor profile, lobe numbers, L/D, shaft diameters, and bearing design and size. The maximum allowable differential pressure is often much less than the maximum discharge pressure. For example, even if the

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II Types of Equipment

TABLE 6.3 Optional Instrumentation for Dry Screw Compressors

Discharge pressure Pressure difference between suction and discharge p2–p1

Pressure ratio p2/p1 Compressor speed for variable speed drivers or turbines Seal liquid temperature (for oilcooled seals or watercooled seals) Seal gas pressure (for all kinds of seals with external gas supply) Seal liquid pressure (for oil-cooled seals or water-cooled seals) Shaft axial thrust

Alarm low

Trip low

In order to prevent excessive overcompression

In order to prevent excessive overcompression

Alarm high

Trip high

If bearings or shafts may be overloaded

If bearings or shafts may be overloaded

Alarm high

Trip high

For compressors with varying operating pressure levels and if discharge temperature may become too high

For compressors with varying operating pressure levels and if discharge temperature may become too high

Alarm high

Trip high

Alarm low

Trip low

Alarm high

Trip high

Alarm low if liquid may freeze or clog

Alarm low if liquid may freeze or clog

Alarm low

Trip low

Alarm high, if seals can be damaged by high pressure

Trip high, if seals can be damaged by high pressure

Alarm low

Trip low

Alarm high, if seals can be damaged by high pressure

Trip high, if seals can be damaged by high pressure

Alarm high

Trip high

design pressure of the casing is 40 bar the maximum allowable differential pressure may be limited to 1.5 MPa by the rotors and bearings. Thus the compressor may be able to compress from 1.5 to 3 MPa but not from 1.5 to 4 MPa. In this case a differential pressure monitoring is required to ensure that the compressor rotors and bearings are not overloaded. Due to the variety of rotor designs and L/D only very rough values for the maximum allowable differential pressure of dry screw compressors can be given. This may range from 0.4 MPa for long rotors with 3/5 profile up to approximately 1.5 MPa for short rotors with 4/6 profile. The allowable differential pressure must be determined on a case-bycase basis. Examples for high discharge pressure or high-pressure difference are no. 3 in Table 6.6 and no. 12 in Table 6.8.

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TABLE 6.4 Standard Instrumentation for Oil-Flooded Screw Compressors Suction Pressure

Alarm Low

Trip Low

Suction pressure

Alarm high

Trip high

To prevent overcompression or high torque Discharge pressure

Alarm high

Trip high

Suction temperature

Alarm high

Trip high

Discharge temperature

Alarm high

Trip high

Thrust bearing temperature (if hydrodynamic)

Alarm high

Trip high

Journal bearing temperature

Alarm high

Trip high

Casing vibration or shaft vibration monitoringa

Alarm high

Trip high

Shaft axial thrust

Alarm high

Trip high

Pressure difference between lube oil and discharge gas

Alarm low

Trip low

Lube oil temperature

Alarm high

Alarm high Trip high

Start-up interlock low Shaft seal oil flowrate

Driveshaft seal oil pressure

Alarm low

Trip low

Alarm high

Trip high

Alarm low

Trip low

Alarm high Temperature difference driveshaft seal oil outlet over inlet

Alarm high

Trip high

a Due to the stiff casing design, 1 accelerometer per compressor or 1 accelerometer at each compressor end (max 2 in total) is sufficient for automatic monitoring with shutdown.

The pressure ratio is limited by the discharge temperature and the undercompression. Extreme undercompression may cause severe gas pulsations in the discharge line. If the pressure ratio is too high the discharge temperature may be higher than allowable for the rotor clearances. The discharge temperature depends also on the gas molecular weight and the isentropic exponent. For dry screw compressors without liquid injection the maximum pressure ratio may range from approximately 2.5 for light gases up to approximately 6 for hydrocarbons with molecular weights of 50 or higher. Again a case-by-case decision is required.

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TABLE 6.5 Optional Instrumentation for Oil-Flooded Screw Compressors Alarm low To prevent excessive overcompression and oil carryover

Trip low

Alarm high

Trip high

If bearings or shafts may be overloaded

If bearings or shafts may be overloaded

Pressure ratio

Alarm high

Trip high

p2/p1

For compressors with varying operating pressure levels and if discharge temperature may become too high

For compressors with varying operating pressure levels and if discharge temperature may become too high

Compressor speed for variable speed drivers or turbines

Alarm high

Trip high

Alarm low

Trip low

Discharge pressure Pressure difference between discharge and suction

With delay to allow normal start-up

p2–p1

Dry screw compressors with liquid injection for discharge temperature limitation can operate with pressure ratios up to approximately 10 in low-pressure applications. A typical example is the styrene monomer process with suction pressure 20–25 kPa abs and discharge pressures of 160–200 kPa abs or even 250 kPa abs (see examples nos. 6 and 7 in Table 6.7). The actual suction flow depends on the compressor size (given by rotor diameter and L/D) and speed. The lower end of the flow is approximately 300 m3/h for the smallest dry screw compressors. The largest dry screw compressors in operation today have an actual volume flow of 77,000 m3/h (see Fig. 6.37, and Table 6.2 nos. 6 and 8, and Table 6.8 no. 14). Dry screw compressors with even larger flows of 120,000 m3/h have been announced but are not yet on the market. The driver power of dry screws can range from <100 kW up to approximately. 6 MW for large single-stage units (Table 6.7, no. 8) and approximately 9 MW for large two-stage machines (Table 6.8 no. 14 and Fig. 6.38). For the ethylene boil off process the suction temperature of dry screw compressors may be as low as 105°C (Table 6.6 no. 2). The lower limit in this case is given by the nil-ductility transition temperature of the materials. Steam compressors have been built with a suction temperature of 100–110°C and a discharge temperature limited to 130–150°C by water injection (Table 6.7 no. 9). The upper

Only for N2-start-up

Oil-cooled mechanical seal

51.7

Liquid injection

Shaft seal type

Molecular weight

Suction temperature (°C)

(bar abs)

Suction pressure

45

1.5

28.02

Fixed speed motor

Driver

(kg/kmol)

Oil-cooled mechanical seal

High molecular weight

Special feature

2103

1.04

–

–

16

35.5

4.08

Double DGS

Fixed speed motor

High suction and discharge pressure, low molecular weight

Hydrogen

3

Fixed speed motor

Low suction temperature

Ethylene boil off

Butadiene

Process

2

1

Nos.

TABLE 6.6 Examples of Single-Stage Dry Screw Compressors Without Liquid Injection

40

1.6

37.1

Carbon ring seal

–

Fixed speed driver

High discharge temperature wet H2S-mix

Hydrogen-sulfide

4

30

21

4.2–28

Continued

Double DGS

–

40%–100%

Variable speed motor

Large range of molecular weight

Hydrogen circulation

5

Screw Compressors Chapter 6

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56 2110

105

13,250

1435

Discharge temperature (°C)

Suction volume flow (m3/h)

Driver rated power (kW)

820

1200

1480

11,400

1230

2550

112

211

104

30.1

6.0

51.5

5

4

3

The bold values are the more extreme condition or range of conditions and operation for these machines.

400

7.0

5.4

Discharge pressure (bar abs)

2

1

Nos.

TABLE 6.6 Examples of Single-Stage Dry Screw Compressors Without Liquid Injection—cont’d

296 SECTION II Types of Equipment

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TABLE 6.7 Examples of Single-Stage Dry Screw Compressors With Liquid Injection 6 Nos.

See Fig. 6.39

7

8

9

Process

Styrene monomer

Styrene monomer

Lime kiln gas

Steam

Special feature

Highpressure ratio, high flow

Highpressure ratio

High flow, high driver power

High suction temperature

Driver

Steam turbine with speed reducing gear

Fixed speed motor

Steam turbine (direct drive)

Variable speed motor 50%–100%

Liquid injection

Demin water

Demin water

Demin water

Demin water

Shaft seal type

Steam buffered carbon ring seal

Double DGS with buffer gas to process

Combined labyrinth/ carbon ring seal

Carbon ring seal

Molecular weight (kg/kmol)

15.5

14.3

38.3

18.02

Suction pressure (bar abs)

0.24

0.27

0.88

1.29

Suction temperature (°C)

36

34

40

107

Discharge pressure (bar abs)

1.9

2.5

4.1

3.47

Discharge temperature (°C)

150

110

90

150

Suction volume flow (m3/h)

77,000

30,000

74,500

10,200

Driver rated power (kW)

2260

1600

5880

690

The bold values are the more extreme condition or range of conditions and operation for these machines.

Fixed speed motor

–

Oil-cooled mechanical seal

47.5/38.5

High molecular weight

Fixed speed motor

–

Oil-cooled mechanical seal

54.2

1.05/2.6

Special feature

Driver

Liquid injection

Shaft seal type

Molecular weight (kg/kmol)

Suction pressure (bar abs)

11

Butadiene

Process

2.0/7.0

Offshore VRU with side stream between stages

Offshore vapor recovery

10

Nos.

7.2/21.5

36–45

Tandem DGS

–

Fixed speed motor

Offshore VRU, high discharge pressure

Offshore vapor recovery

12

1.2/3.5

8–12

Double DGS

–

Steam turbine

High driver power, low molecular weight

Hydrogen recovery

13

1.1/4.4

13.5

Water buffered restrictive ring type

Demin water

Steam turbine

High flow and driver power

Acethylene

See Fig. 6.40

14

TABLE 6.8 Examples of 2 Stage Dry Screw Compressors With or Without Liquid Injection

0.91/2.9

10.9–40.4

N2 buffered carbon ring seal

Demin water

Fixed speed motor

Large range of molecular weight

Flaregas

15

16

1.05/2.8

5.1–23.7

Tandem DGS

Demin water

Variable Speed motor, 50–100%

Large range of molecular weight

Flaregas

298 SECTION II Types of Equipment

1640/500 1150

127/119

6500/4950

3400

83/82

13,400/ 4900

1490

Suction volume flow (m3/h)

Driver rated power (kW)

4450

24,800/8800

189/188

3.4/11.1

34/43

The bold values are the more extreme condition or range of conditions and operation for these machines.

189/188

22.8/32.0

Discharge temperature (°C)

7.7/20

2.7/6.9

64/60

Discharge pressure (bar abs)

32/35

40/40

Suction temperature (°C)

570

4400/ 1340

65,300/ 15,800 8800

105/105

3.4/7.8

70/50

89/97

4.5/11.5

44/50

1000

6700/ 2600

160/160

3.3/7.7

38/38

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FIG. 6.38 Oil-flooded screw compressor package.

FIG. 6.39 Single-stage dry screw with steam turbine.

FIG. 6.40 Two-stage dry screw.

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temperatures are not a fixed limit but the manufacturing clearances at ambient temperature must be much larger than usual to compensate for the thermal growth and the design of the clearances gets difficult. Screw compressors can operate with a large variation of molecular weights as long as the discharge temperature is limited (see Table 6.6 no. 5 and Table 6.8 nos. 15 and 16).

Design Range for Oil-Flooded Screws The maximum discharge pressure for oil-flooded screws is limited by the casing design. Virtually all oil-flooded screw compressors feature a vertically (radially) split casing. Typical design pressures are in the range of 2.5–3 MPa, with some available up to 10 MPa. As with oil-free screws, in many cases the compressor may not be able to operate at the maximum discharge pressure of the casing due to other limitations. The same differentiation between maximum discharge pressure and maximum allowable differential pressure discussed for oil-free screws above also applies to oil-flooded screws. However, with oil-flooded, it is much more feasible for these two values to be the same. For example, it is possible to achieve a pressure ratio of 25:1 in some cases. Undercompression is generally not an issue with oil-flooded screws. This is partly due to the damping effect of the oil on any pulsations in the system. And since the injected oil carries away most of the heat of compression, it is usually possible to avoid thermal limits by adjusting the quantity of oil injection. The actual suction flow depends on the compressor size (given by rotor diameter and L/D) and speed. The lower end of the flow is approximately 200 m3/h for the smallest oil-flooded screw compressors. The largest available today have an actual volume flow of 30,000 m3/h. The driver power of flooded screws can range from <100 kW up to approximately 9 MW for large single-stage units. It should be noted that screw compressor sizes are more often defined by flow than power; therefore, it is uncommon (and technically incorrect) to refer to a screw compressor as, for example, a “2 mW compressor.” Fig. 6.38 shows a photograph of a typical package size and configuration of an oil-flooded screw compressor. From a material standpoint, oil-flooded screw compressors can generally take similar inlet temperatures as oil-free. However, special consideration must be given to the effect of the gas temperature on the injected oil. Very cold or very hot inlet gas temperatures can lead to oil viscosity that is too high or too low. Each case needs to be confirmed by the manufacturer and the oil supplier. Most oil-flooded screw compressors are limited to 120°C discharge temperature. As discussed earlier, it is almost always possible to stay within this limit by injecting more oil.

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Screw compressors can operate with a large variation of molecular weights as indicated in Figs. 6.31 and 6.33. The main concern with high mole weight gas is the tendency of the gas to remain entrained in the oil instead of degassing. Generally, gases with hydrocarbons C6 and heavier should be approached with caution. The oil supplier must confirm the dilution levels and the performance of the oil at the given operating points.

Application Examples for Screw Compressors The examples in Tables 6.6–6.9 are chosen to demonstrate the range and versatility of dry screw compressors.

Rotordynamics Lateral Rotordynamics A detailed evaluation of the lateral rotordynamics of screw compressors are typically not necessary, since compressor manufacturers typically design the compressors per API Standard 619 [1] to avoid lateral rotordynamic issues. In all screw compressors designed per API Standard 619 specifications, the rotor runs subcritical (below the first lateral mode), which avoids synchronous energy (1  running speed) from passing through a critical speed and typically avoids the need for a lateral critical analysis. If it is determined that a screw compressor could potentially be experiencing vibration associated with a lateral rotordynamics issue, the issue should be investigated in a manner generally consistent with other modern rotating equipment. Several notable resources, (such as API619, ISO 10440-1:2007, and API684, among others) provide specific guidance on carrying out such an analysis for machinery of this type. As noted above, the rotor always runs subcritical (below the first lateral mode), which is generally advantageous from an unbalance standpoint as synchronous energy (1  running speed) is not required to pass through a critical speed. Generally speaking, stiffness and damping coefficients should be developed for the bearings based on anticipated loads and geometry. The flexibility of any underlying structural supports should also be considered and accounted for. An undamped critical speed analysis should be performed, along with a damped unbalance response analysis for unbalance distributions that excite the relevant lateral modes. The stability of the system should be investigated using a damped eigenvalue analysis to ensure that whirlrelated problems and instabilities are avoided. In addition, the lateral rotordynamic characteristics of screw compressors are subject to some specific considerations. The lateral critical speeds of many of these compressors are relatively high compared to other rotating equipment, due to the comparatively short and stiff rotors involved. In all screw compressor designs, the rotor runs subcritical (below the first lateral mode), which is

2.02

1.0

Single oil-cooled mechanical seal 16.7

1.3–2.0

86

1000

Varying flow and mole weight, NACE (H2S)

Fixed speed electric motor

Double oil-cooled mechanical seal

15–30

1.3

40

8.4

95

1000–5000

900

Driver

Driveshaft seal type

Molecular weight (kg/kmol)

Suction pressure (bar abs)

Suction temperature (°C)

Discharge pressure (bar abs)

Discharge temperature (°C)

Suction volume flow (m3/h)

Driver rated power (HP)

The bold values are the more extreme condition or range of conditions and operation for these machines.

950

2350

74

20.8

42

Double oil-cooled mechanical seal

Fixed speed electric motor

Low molecular weight

2250

4000–6450

94

18.6

28

2.7

28.0

Single oil-cooled mechanical seal

Fixed speed electric motor

Varying flow

Nitrogen

4

500

3060

85–93

10.1

10–65

1.0

28.2

Double oil-cooled mechanical seal

Fixed speed electric motor

Varying suction temperature

Carbon Monoxide

5

6

300

16.7

21

Fixed speed electric motor

Varying inlet pressure

Hydrogen

Special feature

Fuel Gas Booster

PSA Tailgas

3

Process

2

1

No.

TABLE 6.9 Examples of Single-Stage Oil-Flooded Screw Compressors

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II Types of Equipment

generally advantageous from an unbalance standpoint as synchronous energy (1  running speed) is not required to pass through a critical speed. However, the machines can still respond to typical mechanical excitations (such as misalignment, oil-film instabilities, etc.) and are often subjected to significant energy at the lobe and/or pocket passing frequencies. Care must be taken to ensure that sufficient separation margins are maintained between the predicted lateral modes and prevalent excitation energy over the entire anticipated operating speed range. Another somewhat unique aspect of these designs is that the radial loads developed between screw compressor lobes can generate significant forces, which typically are greater than the rotor weight. As such, these loads should be included, in addition to the gravitational loads, when calculating bearing coefficients.

Torsional Rotordynamics From a torsional behavior standpoint, trains containing screw compressors are similar in many respects to other forms of rotating equipment. Several notable resources are available that outline the requirements for a complete torsional analysis (e.g., see W. Ker Wilson “Practical Solution of Torsional Vibration Problems” [2], and API 684 [3], among others). However, in general a torsional analysis for such a system should include: preparing an appropriate mass elastic model; accounting for all gear speed ratio effects on stiffness, inertia, torque, and stress; determining the system critical speeds; evaluating the mode shapes for potential excitation coupling mechanisms; preparing interference diagrams (for each shaft speed in the train) to determine operating regimes where critical speeds and excitation energy are likely to coincide; and evaluating forced response dynamic stress and torque levels for the entire train during anticipated steady-state operating conditions. The torsional analysis should consider all potential excitation sources, including: lobe and/or pocket passing frequencies from the screw compressors, engine drive order multiples, electric motor slip frequencies, variable frequency drive integer or noninteger excitation components, gear mesh frequencies, and similar sources. In addition, a transient torsional analysis is recommended when synchronous electric motor or variable frequency drives are involved, in order to determine if the machinery can tolerate the stress and torque levels developed during transients such as start-up or motor short-circuit events.

Pulsation and Vibration Screw compressors are positive displacement compressors that inherently generate pulsations (pressure fluctuations). The male rotor typically consists of four lobes; therefore, the main frequency component of the discharge flow is at 4  the compressor speed (4  rpm). Thus, for four lobe compressors, the dominant pulsation excitation frequency is at the primary lobe pass frequency (PPF) of

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4  rpm. Excessive pulsations and vibrations can result in small-bore piping failures. The failures could include cracked lube oil lines (for oil-flooded screw compressors) and instrumentation, which could cause a gas leak and potentially dangerous situation. Therefore, it is important to consider the system pulsations, particularly for higher power compressors. Typically, the relatively standardized silencer designs developed by many compressor manufactures have proven to significantly reduce the compressor-generated pulsations, which results in acceptably low vibrations and noise. Because of the standardized and proven silencer designs and piping acoustic wraps downstream of the compressors, most screw compressor installations do not require a pulsation and/or vibration analysis. In addition to the compressor vibration, pulsation, and noise abatement implementations, the gearboxes are typically placed in a noise enclosure. With these measures many compressor manufacturers can typically achieve a sound pressure level of 85 dB. If the speed, pressure ratio, gas molecular weight, discharge pressure, and over/undercompression are within the manufacturer’s range of experience, a detailed acoustic or pulsation study is typically not necessary. The advice of the manufacturer should be sought. There are some instances when screw compressors can experience pulsation problems. Due to the relatively high running speeds of screw compressors, the pulsation excitation frequencies are typically much higher than those associated with reciprocating compressors. Therefore, one-dimensional (1D) and three-dimensional (3D) acoustic modes can be excited (resonate). Resonance can occur if the compressor frequencies of excitation coincide with the 1D and/or 3D acoustic natural frequencies associated with the pipe and/or vessel geometry and gas properties. However, the shape and location of the acoustic natural frequencies determines whether an acoustic resonance exists or does not exist. Excitation of the acoustic modes (acoustic resonance) can result in piping and/or vessel shell wall vibrations/buzzing, and those vibrations can be amplified significantly if the frequency of an acoustic resonance coincides with the mechanical natural frequencies of the piping and/or vessel shell modes. Pulsation characteristics at lower frequencies can be evaluated with a 1D acoustic analysis. At higher frequencies, acoustic modes associated with the silencer vessel, gas-oil separator/reservoir, screw compressor internal gas-oil passage (IGP), and connecting piping can be excited. These acoustic cross modes (3D modes) can excite shell vibration/mechanical modes involving both circumferential and axial deformation. An acoustic 3D finite element analysis (FEA) can be performed to enable prediction of the acoustic radial or other 3D mode shapes and frequencies associated with the silencers, compressor IGP, and connecting piping to avoid the risk of exciting 3D acoustic and mechanical modes. The 3D acoustic FEA can be coupled with a mechanical FEA of the silencer to predict the significant response frequencies for the silencers, which could lead to high vibration if coincident or nearly coincident with the compressor excitation orders.

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In general, dry screw compressors are more prone to pulsation and vibration issues; however, wet (oil-flooded) screw compressors can sometimes have pulsation and/or vibration issues also. Oil-flooded screw compressors normally do not have suction or discharge silencers, so that there is less pulsation attenuation between the compressor and connecting piping. However, suction silencers are typically less critical, and the oil and “foam” (oil-gas mixture) that is created by adding the oil to the flow path tends to add significant damping to the compressor-generated pulsations. For higher power screw compressors (>370 kW), a Helmholtz resonatortype acoustic filter/silencer (reactive-type pulsation dampener) is typically designed to reduce upstream and/or downstream pulsation amplitudes to acceptable levels. These types of silencers are volume-choke-volume arrangements that are designed such that the acoustic responses associated with the silencer design are below the screw compressor lobe passing frequency (4  running speed) and results in a flow with significantly reduced pulsation amplitudes. This type of design can also result in reduced noise being emitted from the system. However, a reactive silencer that is improperly designed for a given application can result in high shell vibrations and noise amplification. Acoustic filter vessels are most effective at attenuating pulsations at the lower end of the frequency spectrum and can provide some attenuation at the middle and higher frequencies. Passive- or absorptive-type vessels and piping applications can also successfully attenuate pulsations. Absorptive-type vessels are typically good at attenuating the higher frequency content of the pulsations, but these types of vessels are less effective at attenuating the lower and middle pulsation frequencies. Some examples of absorptive vessels are vessels that are lined with or have internals that include sand contained in a relatively thin wall of steel or steel wool restrained with perforated metal sheets and/or steel mesh. Materials restraining the steel mesh can sometimes experience fatigue failure, which could then result in a release of some or all of the absorptive materials. Those materials will then travel downstream of the vessel or fall into the compressor, in the case of top-connected silencers. To attenuate pulsation amplitudes at both lower and higher frequencies, reactive and absorptive designs are sometimes implemented. Absorptive materials can be installed inside volume-choke-volume-type vessels to significantly reduce pulsation amplitudes across the frequency spectrum. Compressor design for the given application can play a significant role in determining whether the resulting pulsation amplitudes are excessive or not. If the Vi does not result in a compression ratio that corresponds with the system application, undercompression or overcompression will occur. Undercompression or overcompression by an excessive amount will result in significantly higher discharge system pulsation amplitudes; however, undercompression is less likely to cause higher pulsation amplitudes.

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If applied properly, orifice plates can be used in screw compressor piping systems to significantly reduce pulsation amplitudes. Typically, an orifice is simply a metal plate with a single hole/bore. Some research has concluded that multi-hole orifices are more effective at reducing pulsation amplitudes at frequencies higher than approximately 100 Hz.

References [1] American Petroleum Institute (API), API 619 Ed. 5/ISO10440-1-2007, API Standard Paragraphs Petroleum, Petrochemical and Natural Gas Industries—Rotary-Type PositiveDisplacement Compressors, 2010. [2] W. Ker Wilson, Practical Solution of Torsional Vibration Problems (With Examples from Marine, Electrical, Aeronautical, and Automobile Engineering Practice), third ed. revised, vols. I–IV, John Wiley & Sons Inc., 1956. [3] American Petroleum Institute (API), API 684 Ed. 2, API Standard Paragraphs Rotordynamic Tutorial: Lateral Critical Speeds, Unbalance Response, Stability, Train Torsionals, and Rotor Balancing, 2010.