Process Plant Machinery--Centrifugal Compressors (C-D)

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SUB-SECTION C: Barrel Compressors

Barrel compressors were pioneered in the early 1970s.

Their specific advantage over high-pressure reciprocating compressors is high delivery volume. Each manufacturer incorporates a combination of design features that are specific to their models of barrel compressor, as well as some features adopted from their regular product line of centrifugal compressors. The operating rotor speeds and pressures involved in each application dictate these selections, as is indicated in the information in this Sub-section.


FIG. C-1 Standard barrel assembly. 1 -Lower half of Inner casing and rotor-resting on the labyrinths- are mounted; 2- Autoclave and front cover bolted to lower casing half, bearings and seals mounted; 3- Mounting of the upper half of the Inner casing and of the piston-rings; 4- Inner assembly or compressor cartridge Is lifted from support and deposited on main and auxiliary rails. Rollers on Inner casing on auxiliary, rollers on front cover on main rails; crane hook removed; 5 - Cartridge pushed Into barrel until Inner casing rollers are supported on the barrel Inside; auxiliary rails removed. Cartridge further pushed into barrel until it’s centering itself; 6 - With erection bolts cartridge Is tightly pulled Into barrel. Front cover bolts are then tightened.

Inner casing assembly support

--- Main railing support; auxiliary railing is supported on main railing

---- Railing i Permanent rollers on Inner casing and removable rollers on front cover


Standard barrel assembly

Compressor Selection

Although the same aerodynamic design and impeller families are used in the axially split R, RZ, RS series and in the radially split RB, RBZ, RBS series, the equipment selection in the field of barrel compressor applications is often complex due to the high pressures and powers involved.


The decision as to whether an axially or radially split casing should be used, depends on: Tightness The radially split design has circular casing joints or flanges with a perfectly even load distribution. The leakage of gas at the two covers can be prevented most effectively. Besides metal-to-metal contact, endless O-rings are inserted in grooves on the two covers. By monitoring the pressure between two adjacent rings the tightness can be controlled. For toxic, flammable and explosive gases the barrel design is therefore always of advantage. For this reason, the API 617 Standard specifies the radially split casing construction for gases containing hydrogen if the hydrogen partial pressure exceeds 13.8 bar (200 psig).

Material stress

The cylindrical design with the smallest possible inner diameter is obviously the most suitable construction. With axially split casings the available space for bolting is moreover restricted at the two shaft penetrations. In order to achieve the required tightness, a high contact pressure at the joints is required. The necessary forces in the bolts are often higher than would be required by the static gas forces if the casing flanges were perfectly rigid and flat. For large compressor frame sizes the radially split design can therefore be the only possible solution even at moderate pressures.

Nozzle location and maintenance

For operating pressures where an axially split design would be perfectly adequate, barrel compressors are often preferred since the nozzles can be arranged in any radial direction. If the necessary space in the axial direction at the non-driven shaft end is available for the horizontal pull-out of the inner casing cartridge, inspections, rotor changes or complete cartridge replacements can be accomplished quickly without removing any process piping to and from the compressor. For tandem units, on the other hand, the first or low-pressure casing should be of the axially split design up to the highest possible operating pressure. A barrel compressor coupled at both shaft ends has to be removed for overhaul or replacement of the rotor. Seals and bearings of Sulzer barrel compressors can, however, always be serviced and replaced with the barrel casing remaining in place.

FIG. C-2 Arrangement of a tandem compressor unit with an axially split LP body and a vertically spilt HP body. ( Sulzer Burckhardt, Switzerland)


The performance of a multistage compressor can be represented in terms of dimensionless coefficients by the following equations:

Dimensionless flow coefficient V Volume flow at inlet (m^3/s, acfm)

u Circumferential or tip speed at impeller outlet (m/s, fps)

Hp Polytropic head (J/kg, Btu/Ib)

He Effective head (J/kg, Btu/Ib)

p~ Impeller polytropic head coefficient n Mechanical speed (rpm)

r/p Polytropic efficiency D Impeller outlet diameter (m, ft)

z Number of impellers

The graph ( Fig. C-3) shows the relationship between the flow coefficient 9 and the efficiency of a multistage compressor. At high ~)-values, the gas velocities in the impeller flow channels increase and lead to a fall in efficiency with increasing ~. At low ~-values, the efficiency falls rapidly because of the increasing influence of the shroud leakage loss and of the unavoidable friction loss in the boundary layers of the impeller disks and blades. It can be seen that the efficiency of a multistage compressor can be improved substantially if the (t)-values of the impellers can be kept in the range 0.03 to 0.09.

If we eliminate u from equations (11C.1) and (11C.2) above we obtain…

For given operating conditions V/v/~p is approximately constant. Although Hp remains nearly the same, the power required is related to the effective head and, as seen in equation (11C.4), depends on the efficiency.

In high-pressure applications, the high-density gases have an intake volume which gives rise to a low (t)-value with correspondingly low efficiency. A larger flow coefficient (t) and hence a more efficient compressor can then only be obtained by:

++Decreasing the diameter, but keeping z and the tip speed u constant, the latter within the admissible stress limit with due regard to the impeller type, method and material of construction. In this case, the mechanical speed increases in direct proportion, since from (11c.3) n ~. u/D. At the high powers required in barrel compressors, the mechanical speed is often restricted from the driver side. If the driver is a turbine, the mechanical speed is often too low for direct coupling with high-pressure centrifugal compressors. The use of a gearbox between compressor and driver gives additional mechanical losses of the order 2 to 2.5%, but the cost can often be recovered with a smaller and more efficient compressor designed for optimum ~. If a gearbox is used, the admissible mechanical speed is again limited for design reasons (pitch line velocity, bending stress on the pinion, bearing speeds and loads).

++Increasing the number of stages without changing the diameter. This results in a speed reduction proportional to Vrz--+--x/~/i where x is the number of added stages. This approach is also limited for reasons related to the design of slender rotors (lower critical speeds).

++A combination of both methods normally allows the best adaption of a compressor to the given mechanical speed of the driver. In practice there are also differences between the head coefficient, axial length and diameter of the different stages that can be used; the polytropic head is not exactly constant, and sophisticated computer methods must be used for the optimization.

The above discussion illustrates an important design consideration for multistage compressors in general, but especially barrel compressors for high-pressure applications.

For a given tip speed or a given mechanical speed, any attempt to design a more efficient compressor by using impellers with higher flow coefficients leads invariably to a more slender rotor. The design of the rotor is therefore of utmost importance, especially in consideration of the fact that modern turbo-machinery must often be designed for operation close to or above the second critical speed.


At lower gas densities, a compressor may operate smoothly at the specified head, volume and speed range. At higher densities and a correspondingly increased shaft input, the compressor rotor may become unstable. This rotor instability manifests itself by unacceptably high subsynchronous vibration amplitudes. Design measures aimed at a greater insensitivity, such as a more rigid rotor, by increasing the shaft diameter and reducing the bearing span or improved damping of the journal bearings, were helpful, but did not eliminate the causes which determine the stability limit of a given rotor or compressor. These became the subject of intensive research and testing programs at Sulzer*. It was demonstrated that rotor instability is caused by the swirling leakage flow around the rotor, especially pronounced at the impeller shaft seals. Swirl brakes mounted at these locations virtually eliminated the causes, known also as cross-coupling forces. The stability limit could be shifted way beyond the highest gas densities as required today for gas reinjection projects, thus solving one of the major problems encountered in modern turbocompressors technology for all practical purposes. Measurements on compressors equipped with and without swirl brakes were moreover in full agreement with theoretical stability limit predictions. Hysteresis of shrink fits or of sleeves on the rotor another potential cause of rotor instability, is eliminated by Sulzer's sleeveless rotor construction and the radial dowel pin impeller fixation.

The results were presented by Sulzer in the following papers:

1. Laterala Vibration Reduction in High-Pressure Centrifugal Compressors, presented and published in the Proceedings of the 9th Turbomachinery Symposium, December 9-11, 1980. Turbomachinery Laboratories, Mechanical Engineering Department, Texas A & M University, College Station, Texas. 2. Prediction of Stiffness and Damping Coefficients for Centrifugal Compressor Shaft Seals, contributed by the Gas Turbine Division of the American Society of Mechanical Engineers and presented at the 29th International Gas Turbine Conference and Exhibit, Amsterdam, June 4-7, 1984. Transactions of the ASME, paper No. 84- GT-86.

Fig. C-4 The chart shows the volume and the corresponding standard Impeller diameters versus the head. The parameter is the number of Impellers. As a measure of the casing frame size Sulzer uses the Impeller diameter. An RB 28-4 is a barrel compressor with a "nominal" Impeller diameter of casing 280 mm and four Impellers. Such a casing can, however, accommodate also impellers up to 315 mm and down to 250 mm in diameter. ( Sulzer Burckhardt, Switzerland)

Application Chart

The chart ( Fig. C-4) is a guide in the first planning phase of projects involving high-pressure compression equipment.

It allows determination of the approximate number of stages required to develop the specified head, the approximate frame size, speeds and powers involved.

The selection chart is calculated for impeller tip speeds not exceeding 260 m/s- this value corresponds to about 80% of the admissible speed for impellers in standard high- tensile strength steels - (t)-values at inlet not exceeding 0.07 and, finally, a scaling down of the impeller diameters in the rear stages. The chart is therefore typical for high- pressure barrel compressor selections where the speed is often restricted for mechanical reasons or due to the use of corrosion resistant materials. The scaling down of the impeller diameter is standard practice in order to increase the ~-values of the last stages and, therefore, the overall efficiency. This aspect is particularly important at high powers.

The mean polytropic stage efficiency would extend from 0.83 for a well-utilized compressor, but designed for a moderate total head or a gas with a low density increase from suction to discharge, down to 0.70 for poorly utilized compressors – inlet ~-values below 0.07 - but designed for maximum head and/or a high density increase from suction to discharge.

The proper selection of barrel compressors requires special skills and expertise in many fields, such as:

++Reynolds number influence at high gas densities

++Mach number influence, especially for gases of high molecular weight

++Properties of gases and of gas mixtures. Different equations of state, such as published by Soave- Redlich- Kwong, Peng Robinson or Starling-Han BWR, may have to be compared.

It is, therefore, recommended that projects requiring high-pressure and high-power compression equipment should be discussed with the manufacturer at an early stage.

Fig. C-5 The Identical elements of R and RB compressor types are: 1 - Rotor with impellers; 2 - Seals; 3 - Bearings; 4 - Return channels; 5 - Diffusors. The different element between the two compressor types is the casing. It will also be noted that the axial casing guide of axially split casings is located on the bearing housings, respectively on the barrel casing of radially split compressors due to the higher forces on the nozzles. * Z is a compressor with interstage "out" and "in" nozzles for external intercooling and liquid separation and/or external mixing with a sidestream or extraction; S is a compressor with an "in" or "out" nozzle only between the respective stage groups for introduction or extraction of a sidestream. ( Sulzer Burckhardt, Switzerland)

Barrel Compressor Design


1 To serve as a pressure containing vessel: Note: Except for special requirements, such as circulators in nuclear reactors, the construction, manufacture and testing of compressor casings is normally exempt from the approval, inspection and certification procedures of the authorities in the state or country of installation. This practice is evidenced, e.g. in the American National ASME Standard, where it’s stated that pressure containers which are an integral part of rotating devices, such as compressors, are not considered to be within the scope of the ASME Boiler and Pressure Vessel Code (Section VIII, Division 1 and 2).

The reference to this code in API 617 implies, however, that casing stress values shall conform to the values recommended for the selected materials of construction at the specified operating conditions.

2 To be gastight: A casing must be technically gastight for compressors handling flammable, explosive or toxic gases.

3 To accommodate and position as mounting base all other compressor components: Diaphragms with return channels and diffusers, shaft seals, bearing housings with bearings, supporting the rotating element.

4 To provide aerodynamically smooth passages: For the acceleration and the deceleration of the gas velocities between suction, discharge and all interstage nozzles of the respective stage or stage groups an optimum aerodynamic design is required.

For axially split compressors these four functions are all met by cast or fabricated casings consisting of an "upper" and "lower" half, bolted together at the "horizontal" joint or dividing flanges.

For reasons of maintenance the nozzles are preferably located in the lower half.

For radially split compressors two concentrically arranged casings accomplish the above four functions, ensuring furthermore maximum ease of assembly and disassembly.

The inner casing is again of the axially split type. Its design entirely fulfills the 3rd function, and together with the outer casing the 4th function. The compressor is assembled in exactly the same way as an axially split type. The design provides unobstructed access for checking all clearances and the correct alignment of the parts. Besides accommodating the nozzles, the outer casing serves exclusively as gastight pressure container for the complete compressor mounted in the inner casing.

Fig. C-6 1 -Inner casing; 2- Cover; 3-Springs; 4-Clearance (variable). ( Sulzer Burckhardt, Switzerland)


The outer casing of a barrel compressor allows optimum utilization of the material and therefore minimum casing weight. It’s of a forged construction with welded-on nozzles, or a casting. The materials of construction are almost exclusively low-alloy steels of high tensile strength.

The inner casing is subject to the gas differential pressures only, with compression being the dominant type of stress. The material of construction is cast steel or nodular cast iron which can be used up to high pressures. For special applications, castings in stainless steel or nonferrous materials can be used.


The bending stress on the diaphragms between the individual stages is proportional to the gas pressure. For high operating pressures the diaphragms are therefore of a reinforced construction, the material being nodular cast iron, cast steel or plate steel. At low pressures cast iron can be used.


The front cover and the autoclave cover are forgings of the same material as the outer casing.

Bearing housings Standard bearing housings for axially and radially split centrifugal compressors are made of nodular cast iron.


During start-up of the cold machine, the inner axially split casing warms up faster than the outer barrel casing. This requires adequate clearances between the two casings and minimum friction resistance. The piston-rings on the inner casing ensure this in the radial direction besides facilitating the pushing in and pulling out of the compressor cartridge. For the required clearance in the axial direction the inner casing is connected to either one of the two covers, the autoclave or the front cover, so that the parts can move axially against each other while having no freedom of movement in any other direction. The inner casing with the two covers mounted thus forms a rigid and self-supporting assembly except for the required axial movement.

On small and medium frames this design requirement is achieved by spring-type connection as shown in the sketch.

The springs are pre-compressed in order to obtain an axially sufficiently rigid cartridge for the final assembly. For large barrels, guiding rods with restricted axial float are used.

Fig. C-7

Casing Design Pressures

On the preceding pages it became evident that the application limits of high-pressure centrifugal compressors are set by the rotating element; i.e. the critical speeds and the stability limit. Beyond these limits the only method to design a reliable compression unit for a given duty is a multiple casing train.

Barrel compressor casings can be built for almost any required final discharge pressure. As an example, Sulzer investigated the feasibility of barrel compressors for 2400 bar (35 000 psia). These compressors would replace hyper-compressors of the reciprocating type. It’s for this reason that the upper design pressure limit of Sulzer multistage axially split centrifugal compressors is of greater interest. At pressures above those applying for these standardized axially split casings, the radially split barrel casing is generally preferred. For special applications and for inert, nonflammable gases axially split casings for higher pressures than those shown on the chart below have been adopted. Examples: An oxygen compressor unit, model RZ 28-6, with a working pressure of 73.50 bar (1070 psia). This machine is the HP compressor of a tandem unit supplying the oxygen for the gasification of heavy residual oils in a refinery.

An air compressor, model RZ271-6, discharging at 62 bar (900 psia). This machine is the HP body of an axial-centrifugal unit for the CAES (Compressed Air Energy Storage) plant at Huntorf, Federal Republic of Germany.

The casings of both the above installations were cast in ductile nodular cast iron GGG 40.

Fig. C-8 Design study of a barrel compressor for a final discharge pressure of 2400 bar (35000 psia). For the rotor an electron-beam-welded construction would be used. Shaft and Impeller hub would be an Integral forging. For the covering disc and the blades the Sulzer standard construction with gold/nickel brazing would be used. 1- Solid electron-beam-welded rotor; 2 - Impeller, Integral with shaft; 3 - Inner casing; 4 - Shrink rings Instead of bolting for Inner casing assembly; 5- Piston seal rings; 6- Front cover; 7- Centering piece; 8- Balance piston stationary labyrinths; 9- Labyrinths; 10- Combined bearing-seals. ( Sulzer Burckhardt, Switzerland).

Fig. C-9 The limits are for information only. The design material stress has been calculated for an assumed casing temperature of 200 degr.

The design pressure of fabricated axially split casings is determined by the actual operating conditions. For medium and large axially split casings segmented case hydrostatic testing can be used. ( Sulzer Burckhardt, Switzerland).

Standard for cast iron axially split casings

Standard for ductile nodular cast iron axially split casings

++,,,-. Standard for low-alloy cast steel axially split casings

--- Maximum standard casing design pressure for forged or cast steel radially split "barrel" compressors

Fig. C-10 Exploded view of barrel compressor showing outer casing, autoclave cover, Inner casing and main cover. ( Sulzer Burckhardt, Switzerland)

Barrel Compressor Design Features Exclusive for RB, RBZ, RBS Series


This standard SULZER design results in an optimum utilization of the material due to equal levels of the hoop stress as well as the stress at the nozzle locations as determined by finite-element stress calculation and measurements. Barrel casings can therefore be designed with minimum weight.

Fig. C-11 Finite-element grids for three-dimensional stress analysis of a barrel casing and its discharge nozzle. ( Sulzer Burckhardt, Switzerland)


TBL. C.1 Centrifugal Compressor Design Features Common for R, RZ, RS and RB, RBZ, RBS Series Barrel

In-line or single-flow Impeller arrangement Sulzer pioneered the in-line impeller arrangement with controlled residual thrust under off-design conditions for high-pressure duties. In 1958, the first high-pressure barrel compressors for 150 bar (2175 psia) and a gas molecular weight of 29 were designed with in-line impeller arrangement. The in-line design produces the lowest alternating thrust at surge conditions and improves the rotor stability.

Controlled Impeller thrust By influencing with small directional blades or other means, the tangential and radial velocity vectors of the gas leakage stream on the outside of the impeller, the static pressure can be controlled. A patent was received in 1970 (Baumann, Eggmann).

Swirl brakes

The same patent was the basis for theoretical and experimental investigations on the rotor stability of the influence of swirl brakes in the gas leakage flow. The exact effects of swirl brakes were measured in the late seventies. The influence of cross-coupling forces could be virtually eliminated. Swirl brakes on part or all stages are now standard at high pressures.

High-efficiency low-flow Impellers Besides the most advanced aerodynamic design the high efficiency of SULZER low-flow impellers is also due to the smoothness of the surfaces, minimizing boundary layer losses. The gold/nickel brazing construction technique, introduced by Sulzer in 1968, is a prerequisite for narrow impellers with smooth surfaces.

Solid quill-shaft thrust bearing The thrust bearing located directly on the coupling shaft is a standard design option for tandem units. It allows the installation of the largest possible thrust area and facilitates maintenance. It also results in minimum bearing spans, lowest critical speeds and minimum overall length of tandem units.

TC thrust collar-type high-capacity turbo-gears Though Sulzer does not manufacture gearboxes, the close cooperation with gear manufacturers allowed Sulzer to take an active part as a consultant in the development of high-capacity thrust collar gears and to develop relevant technical specifications.

Impeller dowel pin The Sulzer standard impeller mounting with a shrink-fit, secured by the symmetrically arranged radial dowel pins, is used exclusively on all multistage compressors since 1974. This design feature allows the use of the largest possible diameters and hence the stiffest shafts. It assures perfect concentricity and balance.

Labyrinths rotating and stationary Sulzer's standard is the use of rotating labyrinths for reasons of safety whenever possible. For very high pressure applications where the gas forces could bend the thin rotating strips, stationary labyrinths are used, especially on the balance piston with the large pressure drop from final discharge pressure to suction. For minimum clearance and leakage losses, special abrasive materials can be used for the stationary portion.



The principal advantage of the Sulzer autoclave cover barrel construction is the complete external assembly of the compressor, including the seals and bearings at both shaft ends.

Alignment, inspection, control and immediate correction of all clearances can indeed be accomplished as if it were an axially split design. This saves time and effort for routine inspections and maintenance work, but also for rotor change-out and for uprating or adjusting the compressor performance map to different operating conditions.


Piston-ring seals instead of O-rings guarantee perfect internal sealing of the assembled compressor during operation. They are not subject to aging. They allow liberal clearances between inner casing and barrel casing. They adapt ideally to the thermal expansion of the inner casing. The friction resistance during assembly is minimal, and the risk of damaging O-rings in grooves is nonexistent.


With bolted covers the barrel casing weight and its length can be kept at a minimum. The bolted cover also enables the complete external cartridge assembly. The inner surface of the barrel remains perfectly even, without grooves for inner rings or seals where O-rings could be distorted or damaged during assembly. The bolted front cover, used up to the highest pressures, also facilitates maintenance as it allows easily accessible flanged connections at the circumference. Removal of high-pressure seal oil and seal gas piping is not necessary in order to gain access to bearings and seals or for cartridge removal.

Shaft seal functions


Compressor seals are of the labyrinth type between the individual stages, on the balance piston and at the shaft ends.

For these noncontact seals the sealant is the process gas.

Its quantity depends on the number of labyrinths, their clearance and the differential pressure between each stage, the balance piston and the suction side and, as regards the leakage to the outside, the suction pressure itself. The internal leakage is continuously recompressed and must therefore be kept as low as possible. At the shaft seals, a process gas leakage can only be prevented by either using a buffer gas or by adding a secondary gastight seal Dry gas seals can be used for flammable or toxic gases. These gastight seals use the process gas or sealant rendering thus a seal oil system superfluous.


Except for air, inert gases and applications at low pressures, gastight seals are standard equipment for toxic, flammable and explosive gases or gas mixtures. The sealant can be furnished by the lube oil system, by a booster system taking the oil from the lube oil system or by a separate sealing system. For applications where any traces of oil in the process gas cannot be tolerated, de-mineralized water or other liquids can be used as sealant.

Separate sealing oil systems are often preferred for the compression of gas mixtures at high suction pressures. They are normally required for any compressor handling gases which by entrainment and dissolution in the oil would reduce the flash point and especially the required minimum viscosity for gearboxes and thrust bearings. This risk could only be avoided by introducing on the labyrinth portion of the shaft seal an inert or a sweet and light buffer gas.

For gas mixtures containing hydrogen sulfide and other acid-forming components, the use of a separate sealing oil system is normal practice. For such cases the contaminated or sour oil returning from the shaft seal can be collected in a separate tank for periodic or permanent reconditioning in an oil reclamation unit.

Fig. C-12 The schematic illustrates the large number of options for compressors equipped with gastight shaft seals.

Increasing costs for gas, oil and energy require close cooperation between client and compressor supplier in order to determine and design the most appropriate sealing system. Special care is also needed for the design, engineering and procurement of the equipment. ( Sulzer Burckhardt, Switzerland)

Since the low-pressure side of the balance piston is always connected to the compressor suction side, both shaft seals operate at roughly equal pressures. In order to prevent any ingress of the sealant liquid into the compressor and hence into the process gas, the preceding labyrinth seal is always purged via a port by a sealing gas stream at a somewhat higher pressure. Unless a buffer gas from an external source is used, this purging or sealing gas is tapped off from the compressor. Most of it flows directly back into the compressor, where it’s recompressed together with the balance piston leakage. A much smaller amount, adjusted to the required minimum velocity in the labyrinth seal by the flow orifices from the seal oil traps and a pressure reducing valve in the feed line, flows towards the gastight liquid-film or mechanical seal, where it mixes with the gas side sealant liquid. This small amount of gas entrained in the sealant is either flared or recuperated in a small gas-oil separator and then returned to the compressor suction.

The supply pressure of the sealant must necessarily be kept above the buffer or purge gas pressure in order to ensure positive sealing and a minimum flow between the stationary and rotating faces of the seal for cooling.

The design of the shaft sealing system depends also on:

++Block-in time

++Coast-down time

++Settling-out pressure and starting pressure

++Sealant supply back-up sources

Centrifugal Compressor Shaft Seal Applications


Labyrinth noncontact seals

Application. Is generally used for inert and nontoxic gases.

It can be used for flammable and toxic gases at lower pressures, provided an inert buffer gas is available (steam, nitrogen). At pressures close to atmospheric pressure, a scavenging system can be used as an alternative or in addition to a buffer gas. Labyrinth seals are also preferred for extreme temperature applications, where a liquid sealant could freeze or evaporate. An application are boil-off compressors in ING re-gasification and storage plants.

Variations. Labyrinth seals can be designed with extremely small, initially negative clearances. Carbon rings and other special abrasive material are then used on the stationary side.

For oxygen service, selected materials of construction with maximum resistance to ignition are used.

Gastight dry gas seal

Application. For flammable, toxic and pure gases up to temperatures of 200 degrees. One single seal is capable of sealing a pressure differential of 100 bar. For higher levels the number of seals is increased accordingly in order to reduce the pressure step by step. Process gas tapped from the compressor discharge is fed to the seal through a fine (2 #) filter. The small leakage is led to the stack. A fan ensures that an ample air flow dilutes at all times any accumulation of gas/oil vapor behind the seal to a safe level, well below an explosive mixture.

Variations. An additional seal serving as standby can be installed (tandem arrangement). Installation of a labyrinth seal with inert gas injection between two dry gas seals is possible.


Gastight liquid-film seal

Application. For flammable and toxic gases. Liquid-film seals are most suited for difficult gases, high pressures and speeds and, in general, all applications where uninterrupted service for several years is required. Their disadvantage is the larger amount of sealant in contact with the gas side of the seal. For sour gas applications, separate sealing oil systems should therefore be employed, unless a sweet buffer gas is available.

Variations. At high pressures the floating ring on the atmospheric side must be of the balanced type. Windback-type floating rings on the gas side permit a substantial reduction of the contaminated sealant return flow. This variation is therefore preferred for gases containing hydrogen sulfide. It requires gravity control of the sealant supply pressure.

Gastight mechanical seal

Gastight double-mechanical seals

Application. The mechanical seal is an option to the liquid-film seal. Due to its advantage (minimum gas side sealant flow) it’s often the preferred seal for clean gases and also for "dirty" gases, except for applications at high pressures (above approximately 100 bar, 1450 psia) and where the required shaft diameter at the seal would result in face velocities in excess of approximately 90 m/s.

Variations. As for liquid-film seals, the floating ring on the atmospheric side must be of the balanced type for high operating pressures. Up to medium pressures the double-mechanical seal is preferred. Besides a lower sealant flow the double-mechanical seal performs as a standstill seal. An overhead run-down tank is not required.

Fig. C-13 Dry, noncontact labyrinth seals with or without ports. Ports are used for discharging the process sealing gas into a low-pressure compressor for recompression, for introducing a buffer gas and/or scavenging. ( Sulzer Burckhardt, Switzerland)

Fig. C-14 Dry gas seal of tandem arrangement. A fan ensures an adequate air flow avoiding any accumulation of gas/oil vapor mixture behind the seal. Dry gas seals require no seal oil system, hence capital costs, energy consumption (no pumps) and requirements are reduced. ( Sulzer Burckhardt, Switzerland).

Atmospheric or air side

Process gas, also sealing or buffer gas

Sealant liquid, seal oil

Sealing gas and sealant liquid mixture: seal gas to flare or back to suction via separator and/or active carbon filter; contaminated sealant to separator or collector tank to waste or to oil reclamation unit.

Gas/oil vapor mixture diluted with air

Fig. C-15 Wet liquid-film seal with floating rings on gas and atmospheric side. For minimum liquid sealant flows towards the gas side, the gas side rings can be of the wind back type (upper half), in order to reduce the "sour" seal oil quantity. The threaded windback grooves act as the stator of a pump. The atmospheric side ring can be of the "unbalanced" or of the hydraulically balanced type (lower half). The shaft at the seals is protected by chromium plating or by replaceable sleeves. ( Sulzer Burckhardt, Switzerland)

Fig. C-16 Wet single-mechanical (contact) seal. On the atmospheric side the seal is of the liquid-film type. The floating ring can be of the "unbalanced" or of the hydraulically balanced type (lower half). R Reference pressure (connected to the seal at the compressor discharge side)

Sealant overflow for cooling of seal. ( Sulzer Burckhardt, Switzerland)

FIG. C-17 Wet double-mechanical seal. This seal requires the least quantity of sealant liquid. Most of the sealant, entering below, Is used for cooling and Is therefore flowing directly back to the reservoir. The application of the double-mechanical seal Is, however, restricted to low and medium pressures. ( Sulzer Burckhardt, Switzerland)

Couplings, Bearings For Barrel Compressor


The low density of gases in comparison to liquids requires compressors and turbines operating at the highest possible mechanical speeds. Couplings must be carefully selected with due regard to

++normal, transient and especially pulsating torques;

++movements of the coupled rotors and their casings due to thermal expansion and other causes;

++maximum combined axial thrust of each component under extreme conditions, such as e.g. surging or loss of vacuum in case of condensing turbine drivers.

At the high speeds and powers dry, nonlubricated couplings of the quill-shaft or of the metallic-diaphragm type are generally preferred to all other coupling types. The requirements for low weight and high balance quality often exclude the use of other coupling types. Toothed-type coupling depend moreover on the steady supply of lube oil, free from solid particles. Another element requiring careful selection and dimensioning are the thrust bearings. Whenever possible, the total number of thrust bearings per shaft string should be kept at a minimum. For this reason the order of preference is:

++Solid couplings with a flexible quill-shaft as per API1671, 1st edition, 1979, paragraph 2.1. The quill-shaft allows the elimination of high-speed thrust bearings of geared units and, for direct driven units, the use of one single-thrust bearing.

++Diaphragm-type couplings as per API671, 1st edition, 1979, paragraph 1.1.3. Diaphragm couplings are necessary when the two connected rotors or gearbox shafts are or must be equipped with their own thrust bearing. The diaphragm acts as a soft spring in the axial direction. Besides the required misalignment capability, it must be designed for compensating the maximum combined axial displacement of the two coupled shafts from their fixed points.

Both coupling types should also be given preference at lower speeds, where other coupling types may, however, offer special advantages. For motor drivers, toothed-type couplings with limited axial end float and, in special cases, maximum torque limiting shear bolts can be used on the motor side besides couplings with transient and oscillating torque dampening capabilities.

The hubs for high-speed couplings should be tapered and designed for hydraulic fitting without keys. At low speeds and powers, straight bores and keyed shrink tiffing can be used.

Fig. C-18 Solid quill-shaft coupling. ( Sulzer Burckhardt, Switzerland)

Fig. C-19 Tilting pad self-leveling or self-equalizing thrust bearing with Individual oil Injection for each pad for flexible shafts. For stiff shafts, such as gearbox shafts, self-leveling pad support is not required. ( Sulzer Burckhardt, Switzerland)

Fig. C-20 Two-lobe radial bearing, type R 2F, for heavy rotors and speeds normally below 7000 rpm. ( Sulzer Burckhardt, Switzerland)

Fig. C-21 Four-tilting-pad radial bearing, type R 4K, for high speeds and normally used in barrel compressors.( Sulzer Burckhardt, Switzerland)

Fig. C-22 The main and intermediate gearboxes of the below-shown compressor unit are of the single-helical thrust collar type, allowing the use of solid quill-shaft couplings with one common low-speed thrust bearing on the main gearbox. This compressor unit serves a nitrogen reinjection scheme for enhanced oil recovery (EOR) in the United States. The nitrogen is compressed from 2.46 to 58.00 bar (35.7 to 840 psia) injection pressure. For injection pressures above approximately 80 bar (1160 psia), the HP casing would be a radially split barrel compressor. The compressor casings are cast in ductile cast iron. ( Sulzer Burckhardt, Switzerland)

Shaft String Layout

Fig. C-23 Shaft string layout. ( Sulzer Burckhardt, Switzerland)


~ Compressor Layouts 1 to 8 show tandem units. If the second or HP compressor only is of the radially split design, it should be located at the end for axial pull-out of the cartridge. Layouts with a single compressor are similar as indicated by the dotted line. The thrust bearing on the gear of layouts 3 and 4 would then be on the direct driven compressor. Layout 2 shows the preferred thrust bearing location on the intermediate quill-shaft for tandem compressors operating at the same speed.

Driver The arrow indicates the required sense of rotation. R and RB compressors always rotate clockwise in the direction of the gas flow.

TC (Thrust Collar) type gearbox If single-helical gearboxes without thrust collar are used, layouts 3, 4, 5 and 6 would require one, layouts 7 and 8 two additional thrust bearings on the high-speed side.

H Solid quill-shaft coupling H Diaphragm coupling A low speeds other coupling types with axial displacement capability could be used, often designed with limited actual float for electric motor drivers, equipped with journal bearings only.

Thrust bearing, radial bearing

Layout 9 shows a double-helical speed increaser. The arrow shows an alternate thrust bearing location on the gearbox.

With single-helical gearboxes, the layout would be similar except that the gearwheel would also require a thrust bearing.

Layout 10 shows an epicyclic gearbox allowing the same or opposite rotation of the output shaft. Layout 11 is an arrangement with a multiple-shaft gearbox with two outputs. This layout can be of advantage at high speed ratios of multiple-casing barrel compression units and where space is restricted.

Materials of Construction For Barrel Compressor

The gas or gas mixture to be compressed and the required physical properties determine the most important selection of the materials of construction. Besides the recommendations of industry standards and of the suppliers of the materials, the experience of the user and of the compressor manufacturer should be taken into consideration. This is especially important for corrosive and wet gas service. An important example in this regard are gases containing hydrogen sulfide. In such cases the materials of construction should conform to the NACE Standard MR-01-15 and subsequent revisions.

The materials of SULZER compressors as tabulated below are limited to the most frequently used specifications and the most important components: impellers, shafts, SULZER solid couplings and the casings.

Quite often materials of different qualities can be proposed.

Examples are casings in ductile nodular cast iron as an alternative to steel. In other cases both stainless and nonstainless materials are to be considered. For the impellers, martensitic stainless steels are often preferred since they offer a higher strength, allowing somewhat higher impeller tip speeds. In such cases the most appropriate selection should be reviewed carefully and jointly by the user and the equipment manufacturer.

The certification that the ultimately selected materials conform to their specifications is of utmost importance. The quality assurance measures and methods of testing should be clearly defined, and all parts for which certificates, witnessed or observed tests are required should be listed. It’s Sulzer's practice to submit, together with the compressor quotation, defined quality assurance and testing programs for all vital parts and to comment explicitly on the deviations, if any, from the requirements of the material requisition.

The governing material specification is the one in general use in the country of the material supplier. For SULZER compressors produced in Europe, forgings and castings are normally procured from European suppliers.

The temperature limits for which the materials as listed below can be used, depend on the actual requirements as regards tensile and impact strength. Materials for very low temperatures have not been listed. (Material, TBL. C.2 on facing page).

Barrel Compressor Applications

Applications for barrel compressors have grown very significantly since their introduction, as is indicated by the figures below.


The fact that oil and natural gas are precious, nonrenewable resources contributes to a diversified and fast-growing market for centrifugal compressors of both the axially and, especially, the radially split barrel type. The selection of the most suitable compression equipment and the drivers depends on the conditions, differing from site to site, and the resultant technical and economical considerations, such as

++the type and geographic location of the field, the products--oil, commercial natural gas, LPG, NGL, LNG- as well as the markets and the mode of transportation which are in turn related to

++the methods of production: reinjection of gas for conservation and for maintaining reservoir pressure, EOR recovery with inert gases, gaslift operation, insitu combustion.

Successful operation of a barrel design for the compression of pure oxygen up to 140 bar (2030 psia) was a major step toward high-pressure oxygen plants, as required for the production of synfuels based on coal conversion.

For cracking and upgrading processes in refineries, barrel compressors for hydrogen-rich gases over a wide range of pressures and flows are required. An impeller layout aimed at a wide stable operating range is often mandatory. The barrel-type centrifugal compressor competes with positive displacement-type compressors. For the latter, standby units are normally required. It’s for this reason that a single barrel compressor unit is often preferred.

For the compression of the light synthesis gas mixtures in tonnage methanol and ammonia plants, multi-casing barrel compressor trains are standard equipment.


TBL. C.2 Part Sulzer/DIN specification form Comparable or equivalent U S specification Notes

++For welded impellers

++Good resistance against stress cracking corrosion (sour gas service) if heat-treated to a maximum yield strength not exceeding 620 N/mm 2 (90 000 psi). Admissible rated impeller tip speed is then limited to approx. 285 m/s, depending on wheel type

++For welded impeller constructions

++Standard material for high tip speeds up to 360 m/s, depending on impeller type and heat treatment

++For welded and gold/nickel vacuum-brazed impellers

++Impeller material if stainless steel is required

++NACE (editorial revision 1984) accepts this material with 23 HRC

++For welded and gold/nickel vacuum-brazed impeller constructions

++NACE (editorial revision 1984) accepts this material with 31 HRC which allows substantially higher tip speeds for corrosive gas service


Fig. C-24 Aerial view of the El Morgan complex in the Gulf of Suez. The platform on the right shows the gas-lift compression modules engineered and furnished by Sulzer. The weight of each module is approximately 250 t. The five identical compressor units are driven by SULZER type $3 gas turbines. After a total of over 200 000 accumulated running hours at the end of 1984, the average utilization factor was 91.05%, the availability factor 96.3% and the reliability factor 99.3%. ( Sulzer Burckhardt, Switzerland)

Fig. C-25 Reinjection compressor unit for the Moomba Field in Australia. The tandem barrel compressor unit is designed for a hydrocarbon gas mixture with a molecular weight of 32 and a final discharge pressure of 202 bar (2930 psia). The driver is an industrial gas turbine with a site power rating of 2890 kW and an Integrated gearbox.

( Sulzer Burckhardt, Switzerland)

Fig. C-26 Syngas compressor unit for an ammonia plant during shop assembly. The recycle compressor is on the left, the feed-gas compressor on the right. The discharge pressures are 227 and 236 bar (3300 and 3450 psia) respectively.

For this plant, a 13 000-kW electric motor driver had been chosen. The required two-stage gearbox, rated 13 MW, with a total speed ratio of 1490/14 530 rpm also provides the driving shafts for the mechanical lube and seal oil pumps. ( Sulzer Burckhardt, Switzerland)

Fig. C-27 Axially split medium- and radially split high-pressure compressor of a gas Injection compressor train for an enhanced oil recovery (EOR) project in Texas, USA. The plant consists of three injection trains with a total Installed power of 45 000 kW. The low-pressure compressor, model RZ 71-7, is coupled to the steam turbine driver. The medium- and high-pressure compressors as shown below, models RZ 35-8 and RB 28-7, are driven via an intermediate gearbox of the thrust collar type from the low-pressure compressor. The gas is compressed from 1.0 to 130 bar (14.5 to 1885 psia) Injection pressure.

Fig. C-28 Offshore modules for the Bombay High South oil and gas field. The modules with a weight of approximately 1300 t were engineered by IHI and completed in the IHI Aichi works In Japan In December 1981. The modules are for associated gas service and are equipped with tandem barrel compressor units, type RBZ 45-6 plus RB 31-7, driven via thrust-collar-type gearbox by an aero-derivative gas turbine with an ISO rating of 14780 kW. The compressors were built by the SULZER licensee Ishikawa-Jima-Harima Heavy Industries Co. Ltd. (IHI). Each unit compresses 2.4 MMNm3/d gas from 7.5 to 87.8 bar (109 to 1270 psia). ( Sulzer Burckhardt, Switzerland)

Fig. C-29 One of three identical units for an offshore gaslift operation in the Persian Gulf. The tandem units are driven via gearbox by a 10 360 kW industrial gas turbine.

The gas is compressed from 13 to 152 bar (188 to 2200 psia). The compressor casings, model R 28-7 for the LP duty and RB 28-8 for the HP duty, are equipped with mechanical seals. The barrel compressor, located at the end of the shaft string, is shown with the Inner cartridge removed for maintenance. ( Sulzer Burckhardt, Switzerland)

Fig. C-30 High-pressure reinjection unit for the Airar Field in the Sahara. The three identical units compress natural gas from 81.7 to 322.7 bar (1185 to 4680 psia) and are driven by industrial-type gas turbines. For the work tests at full load, nitrogen was used up to a discharge pressure of 388 bar (5630 psia) in order to verify the rotor stability. The gearboxes were furnished by Maag. They are of the thrust-collar design with a throughput of 21 830 kW at the rated speed ratio of 4670/12 924 rpm. ( Sulzer Burckhardt, Switzerland)

Fig. C-31 Tandem unit for gas compression duty on a platform in the North Sea. For this plant three identical units were furnished, mounted on three-point support skids with motor and gearbox. The lube and seal oil systems are located on separate skids. The picture shows the two compressors with acoustic enclosures, ready for shipment. ( Sulzer Burckhardt, Switzerland)

Fig. C-32 Syngas compressor train in a 1000t/d methanol plant. The unit Is driven by a high-efficiency, "low-speed" (8000 rpm) condensing-extraction turbine. The four barrels, operating in series, are arranged in a H-type configuration on the two sides of the multiple shaft gear-box. Compressors and the single driver can therefore be designed for optimum speeds. The resultant gain in efficiency Improves the overall energy consumption besides compensating for the relatively large mechanical losses of the 25 MW gearbox. ( Sulzer Burckhardt, Switzerland)

Fig. C-33 For gas transmission projects, multistage barrel compressors in pipeline configuration - horizontally opposed nozzles - are used, allowing direct coupling and optimum efficiency. The picture shows the Ruswil station, equipped with a Sulzer barrel compressor, driven by a Sulzer split-shaft gas turbine, type $3, with four combustors. Sulzer gas turbines are used for crude oil and gas pipelines in Romania, Iran, Egypt, Libya, Germany, and Switzerland. ( Sulzer Burckhardt, Switzerland)

Fig. C-34 Refinery gas compressor, type RB 31-5, on the Sulzer ( Zurich) compressor test stand. For this performance test, the job lube and seal oil systems as well as the test stand gearbox and motor were used. The compressor is designed for compressing a hydrogen-rich gas with a molecular weight of 4.7 from 156 to 188 bar (2260 to 2730 psia). The job driver as a Siemens high-speed condensing turbine. After the performance test, a string test with the job turbine at full speed and power was conducted on the Siemens turbine test stand in Wesel. ( Sulzer Burckhardt, Switzerland)




Isotherm Turbocompressors

Fig. D-1 Air compressors, type RIK 80 and RIK 56. Transportation as a single-lift package. (Sulzer Turbo Ltd, Zurich, Switzerland)

Common Features

Common Features are incorporated to best satisfy the requirements of the end-user, such as: 1 Low power consumption 2 Resistance against corrosion 3 High rotor stability - low vibration level 4 Low noise level 5 Simple shaft-string configuration 6 Easy erection 7 Simple maintenance 8 Minimum space requirement 9 High reliability And this is how it has been achieved and proved in hundreds of industrial applications: 1 Low power consumption

++Intensive intercooling of the gas reduces the inlet temperature into the subsequent stage and therefore its power requirement.

++Optimization of the distribution of the total cooling surface within the individual cooling stages with respect to heat load, cooling effect and air-side pressure drop adds to the high overall efficiency.

++The skillfully designed short flow path achieved by the single-shaft in-line design with cooler tube bundles integrated in the casing, reduces the pressure losses on the gas side as there is no external piping.

++The staggered high-flow impellers ensure optimal combined performance of all stages, avoiding the lower range of flow coefficients which exhibit a progressing sharp drop of efficiency ( Fig. D-2). All impellers are of fully welded or welded and brazed construction with the blades shaped in three dimensions ( Fig. D-3).

++In the case of very dirty cooling water, an automatic cleaning system for the cooler tubes can be installed.

TBL. D.1 Type Range

Fig. D-2 Influence of flow coefficient of the first Impeller on the efficiency of subsequent stages. (Sulzer Turbo Ltd)

Fig. D-3 Welded Impeller of high flow coefficient.

(Sulzer Turbo Ltd, Switzerland)

Fig. D-4 Temperature and humidity conditions: m hot a cool, but well above dew point; M cold, but not yet saturated

++cold and condensing (Sulzer Turbo Ltd, Zurich, Switzerland)


TBL. D.2


1 Solid, sturdy rotor with shrunk-on dummy piston 2 Shrink fit secured by symmetrically arranged radial dowels for impellers 3 No shaft sleeves between stages 4 Labyrinths always on the rotating element.

Stainless steel strips caulked into the shaft and impeller grooves 5 Nickel plating or other coating of shaft portions exposed to corrosion, if necessary 6 Tilting-pad radial bearings for higher speeds 7 Solid coupling, tightly bolted to flexible intermediate shaft


Minimum sensitivity to critical speeds and unbalance due to higher rotor stability; reduction of rotor internal damping No need for keys and distance bushings; fixation assures concentricity and perfect balance under extreme operating conditions; allows larger shaft diameters; reduces stress on shaft and impeller Reduces rotor hysteresis and increases running stability No distortion of rotor due to local heating-up in case of rubbing; labyrinths can be refitted easily Plating instead of shaft sleeves is a more direct protection; allows larger shaft diameters Improves running stability; no oil-whip; higher external damping Improves reliability due to elimination of high-speed thrust bearing and toothed-type couplings; no torque lock thrust on high-speed thrust bearing


This would extend time between overhauls without impairing long term efficiency.

2 Resistance against corrosion

++Most of the carefully designed flow path is handling hot superheated air;, the not quite saturated air after the cooler is taken by the shortest raise to the next impeller ( Fig. D-4).

++Thanks to the vertical position of the built-in lateral cooler tube bundles the inertia-type water separators fitted at the outlet of the coolers (except for the RIO types) show a very high separation efficiency enhanced by the effective condensate removal by mere gravity ( Fig. D-5 and 6). Due to this and owing to the sub-cooling effect along the tube fins, the air entering the following stage has a mean temperature just slightly above the dew point, which again inherently reduces the proneness to erosion and corrosion.

The condensate is drained by automatic traps.

3 Rotor stability

In conjunction with the radial bearings and the special coupling technique, turbocompressor rotors are designed for inherently high stability under all practical operating conditions. This is achieved by the following main features ( Fig. D-7):

4 Low noise level

The sturdy radial casing with the built-in coolers have an attenuating effect on the noise generated by the active high-velocity parts embedded in this compact outer package. The same applies to the double-casing axial part of the ARI series.

The noise level is therefore lower than that of a centrifugal compressor with separate external coolers and the necessary interconnecting piping. In case of severe noise level restrictions, the compactness of the compressor results in a simple and inexpensive noise hood covering compressor and gear.

5 Simple shaft-string configuration

The single-shaft in-line concept allows a very simple configuration of a complete motor or steam turbine driven compressor train with a minimum of shafts and bearings. The common solutions generally are:

a Steam turbine direct drive with one common axial thrust bearing in the steam turbine and solid coupling with flexible intermediate shaft between turbine and compressor. Four journal bearings. Axial thrust compensated. No additional load on thrust bearing due to torque lock. Standard for ARI and semi-packaged RIK types.

b Steam turbine direct drive with individual thrust bearings.

Four journal bearings. Axial thrust not compensated.

Additional load on thrust bearing due to torque lock caused by thermal expansion of shafts taken up by conventional gear coupling. Depending on magnitude of transient thermal expansion of compressor and turbine rotor, a diaphragm-type coupling can be used, reducing the additional axial thrust and requiring no lubrication. Alternative for ARI and semi-packaged RIK types.

c Motor drive with speed increasing gear of conventional design. Axial thrust bearing on compressor and gearwheel shaft. Eight journal bearings including gear and motor.

Solid coupling and flexible shaft between compressor and gear; no torque lock. Solution suitable for high-power low- speed compressors. Preferred alternative for the large ARI types.

d For higher speeds with motor drive, the technically and economically superior solution, described in Fig. D-9 to 11D-11 is applied.

Fig. D-5 Intercooler outlet with water separators. (Sulzer Turbo Ltd, Zurich, Switzerland)


Particularly for electric motor driven high-speed compressors of medium to high power and/or pressure, it’s normal practice to make extensive use of solid couplings allowing the use of only one axial thrust bearing for single- or multiple casing arrangements.

An intermediate shaft, flexible enough to allow for considerable misalignment, is inserted between the two shaft ends of the machines to be coupled together ( Fig. D-9). In the case of motor-driven units, the normal technique is to use single helical gears provided with thrust collars on the pinion shaft, as shown in Figures 11D-10 and 11. The thrust collars not only neutralize the axial thrust created by the meshing of the teeth cut at an angle to the axis of the shaft, but also transmit the residual axial thrust of the high-speed rotor train to the thrust bearing on the low-speed wheel shaft.

Good gear meshing requires “parallelity” of gear and pinion shaft and automatically assures parallelity of the contact surfaces of thrust collar and wheel rim. The slight tapering of the thrust collars is responsible for the formation of a wedge-type oil film creating a pressure zone spread out on an enlarge surface with a pressure distribution very similar to that of a standard oil-lubricated journal bearing.

The relative motion between the two contact surfaces of the thrust collar system is a combination of rolling and sliding and takes place near the pitch circle diameter, resulting in a very small relative velocity. The thrust transmission is therefore effected with almost no mechanical losses. The considerably reduced losses of the single thrust bearing on the low-speed shaft as compared with the high losses of individual thrust bearings on the high-speed train lead to a substantial power saving. Moreover, this low-speed bearing can be more amply dimensioned to provide a much higher overload capacity.

This coupling arrangement avoids heavy overhung gear couplings which are usually responsible for not clearly defined lower critical speeds and for the phenomena of torque lock leading to additional loading of the axial thrust bearing. The resulting axial friction forces can become quite substantial if insufficient attention is given to the cleanliness of the lubricating oil. This arrangement is, therefore, the preferred solution.

Its strict application is clearly visible on the air compressor train ( Fig. D-8).

6 Easy erection on site and dismantling for inspection due to:

- Package construction

- One single horizontal plane of the axis

- Vertical cooler bundles easily withdrawable

- No heavy and cumbersome cross-over piping between compressor and external intercoolers

7 Simple and low-cost maintenance, because all components are easily accessible.

8 Minimum space requirement through compact single-shaft design with integrated coolers. Low elevation of operating floor for ARI types; skid-mounted single-lift package with integrated gear and lube oil system for RIK and RIO types.

9 High reliability by using well-known elements proven in hundreds of machines of the same basic design.

Fig. D-6 RIK model with intercooler tube bundles in the casing bottom half. The vertical water separators at the cooler outlet ensure effective drainage of the condensate. (Sulzer Turbo Ltd, Zurich, Switzerland)


++Two-lobe journal bearings are used on the larger frame sizes of the ARI series running at a moderate speed ( Fig. D-12).

++Tilting-pad journal bearings are incorporated in the RIK and RIO series operating in a higher speed range. They contribute to the high rotor stability at high rotational speeds ( Fig. D-13). The horizontally split journal bearings are white-metal-lined and forced-feed-lubricated. Adjusting plates with a slight curvature in axial direction allow the bearings to be set accurately on erection. Shims placed between the plates and the bearing shell make a possible corrective realignment very easy.

Thermoelement connections for white metal temperature measurement are fitted.

++The axial thrust bearing is normally located on the low-speed shaft of the gear. In multicasing arrangements with no gears it’s normally located on the intermediate shaft. The thrust bearing is fitted with a load equalizing system. The pads are individually lubricated ( Fig. D-14). NP = reference point (100 %)

= design point

(x = angular position of the inlet guide vanes (RIK models) or the adjustable stator blades (ARI models) Valid for air at constant inlet data.

Depending on the specific process requirements, such as higher over-load capacity, a certain pressure rise to surge, maximum efficiency at design point or rather at a certain part load, the process design point NP may be placed differently in the characteristic curve.

Fig. D-7 Design principles of Sulzer turbocompressor rotors. (Sulzer Turbo Ltd, Zurich, Switzerland)

Design Features (RIK)


The various frame sizes are skid-mounted units with built-in intercoolers and integrated lube oil system. They are suitable for mounting at grade. The erection on site is reduced to mere setting of the skid on a simple foundation slab, lining up with driver and connecting gas and cooling-water piping ( Fig. D-22). Alternatives to the standard motor-driven concept are possible.

For example:

++Compressor directly coupled to the steam turbine driver or expander

++A booster coupled to the compressor, with or without intermediate gear

++Separate freestanding lube oil system

++Suction nozzle facing downwards


The horizontally split casing contains the five centrifugal stages and three pairs of vertical intercoolers ( Fig. D-23). The axial inlet ensures ideal flow conditions through the inlet guide vanes into the first stage. The bearings can be inspected without having to disconnect the gas or oil piping nor to disturb the casing top half. The efficient aerodynamic form of the flow passages to and from the coolers is the result of exhaustive model tests and ensures that the gas flow in each stage is equally distributed between the parallel cooler elements ( Fig. D-24). For the hydrostatic tests the casing is divided into several chambers and submitted to a water pressure of 1.5 times the maximum possible operating pressure of the corresponding compartment.


The shaft and interstage seals are of the labyrinth type. The stainless steel strips are fixed in grooves in the rotating parts (shaft, impeller hub and cover disc) and have a very small radial clearance against stationary plastic ring segments in the corresponding partition walls. Internal and external leaks are thus kept to a minimum ( Fig. D-7).

Fig. D-8 Typical shaft string configuration of a motor-driven ISOTHERM compressor with booster. Axial thrust transmission according to Figures 11D- 9 to 11 with one single thrust bearing on the low-speed side of main gear. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-9

Fig. D-10 Method of axial thrust transfer in a single helical gear with thrust collar Fu - Peripheral force; FA - Axial force; u - Peripheral speed; p - Pressure. (Sulzer Turbo Ltd, Zurich, Switzerland)

The axial thrust of the impellers is almost entirely compensated by a balance piston at the discharge end of the compressor.

The piston is provided with labyrinth strips rotating against a white-metal-lined steel ring.

With the in-line arrangement of the impellers the resulting total axial thrust of the rotor is, contrary to a back-to-back arrangement, always acting in the same direction and of the same order of magnitude under all operating conditions (normal, reduced load, surge and rotating stall). The balance piston is dimensioned in such a way that the compensated residual thrust is reduced to a minimum, but still always acting in the same direction. With this method the axial thrust bearing need not to be oversized, and the bearing losses are reduced accordingly without any risk of overloading it under abnormal operating conditions.

On the suction side, a special sealing system prevents any oil or oil mist of the bearing space seeping into the surrounding suction ducts thus contaminating, For example, the air of an oxygen or nitrogen plant ( Fig. D-25) The sealing air introduced in the middle of the shaft seal is discharged to atmosphere on the bearing side in order to avoid building up of pressure in the confined bearing space connected with the oil tank.


The advantages of the rotor design of Sulzer turbocompressors were described in detail on Fig. D-7. The impeller and shaft materials undergo a number of metallurgical tests.

Further tests are carried out during manufacture. The finished impellers are then balanced at low speed, subjected to an overspeed test. The assembled rotor is dynamically balanced over the whole speed range up to full speed. Impellers with a medium to high flow coefficient are of the fully welded construction with the blades shaped in three dimensions (see Fig. D-3). Small and narrow impellers are of the combined welded/brazed construction. The sense of rotation of the rotor is clockwise, seen from the suction side.

Fig. D-11 Transfer of external forces. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-12 Two-lobe journal bearing. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-13 Multi-segment journal bearing with four tilting pads. (Sulzer Turbo Ltd, Zurich, Switzerland)


To obtain the characteristics as shown in Fig. D-18 with an infinite number of operating points between maximum performance and surge line, inlet guide vanes are fitted ahead of the first stage. They are actuated by a pneumatic servomotor which may be connected to an automatic pressure or flow controller ( Fig. D-27). The guide vanes are pivoting in self-lubricating bushes and are connected to an adjusting ring by a maintenance-free linkage system ( Fig. D-28). Another link connects the ring with the pneumatic actuator.

The absence of a lubricant avoids contamination of the process gas. When starting the compressor, the guide vanes are in an interlocked closed position to reduce the starting torque to a minimum.


Three pairs of intercooler tube bundles are mounted in a vertical position on each side of the casing and bolted to the water box. They can expand freely. Round finned tubes ensure excellent heat transfer on the air side. The gap between tube bundle and casing at the exit of the cooler is sealed with a rubber membrane to avoid bypassing of uncooled air.

Design Features ( RIO)


++Compact five-stage centrifugal inline design

++Three pairs of vertical cooler bundles integrated in casing

++Nominal discharge pressure up to 20 bar

++Directly coupled booster compressor available for high discharge pressure


Compressor series lies in its advanced design and the thorough cooling of the medium by built-in intercoolers. The compression approaches the ideal of efficient isothermal compression. Result: lowest possible energy consumption and a very compact machine. There are no external coolers, no crossover piping, no expansion joints. Its simple, lightweight package requires less space, has low overall profile, is easy to erect and results in minimum installation cost.

Fig. D-14 Kingsbury-type axial thrust bearing with self-equalized pads with directed lubrication. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-15 Determination of the absolute humidity x and the molecular mass Mf of the wet air. (Sulzer Turbo Ltd, Zurich, Switzerland)


TBL. D.3 Performance Data RIK and ARI

Type designation Compressor size Operating condition Indices Nominal diameter Power input Mass flow Suction pressure Suction temperature Relative humidity of the air or gas Discharge pressure Molecular mass The following factors and symbols are also used for the calculation: Suction volume (actual)

Absolute humidity Isothermal efficiency Suction branch Discharge branch Dry Wet D (cm)


TBL. D.3 (continued) Selection and performance calculation of an ISOTHERM compressor

Fig. D-16 Determination of the discharge temperature. (Sulzer Turbo Ltd, Zurich, Switzerland)

FIG. D-17 Compressor selection diagram. (Sulzer Turbo Ltd, Zurich, Switzerland)

FIG. D-18 RIK compressor with Inlet guide vanes and constant speed driver. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-19 RIK compressor without inlet guide vanes, but running at variable speed. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-20 ARI compressor with adjustable stator blades and constant speed driver. (Sulzer Turbo Ltd, Zurich, Switzerland)

Design Features ARI


Where the ideal concept of the centrifugal ISOTHERM compressor reaches the economical limits with regard to size, weight and specific cost, the axial compressor design offers an attractive and technically convincing solution for large volume flows.

For equal aerodynamic loading (Mach number) of the machine and for the same tip diameter of the rotor, an axial stage will handle a volume flow about twice as large as the one of a wide centrifugal impeller. If, therefore, the centrifugal section of an isotherm compressor is preceded by an axial booster with a pressure ratio of about 2, the centrifugal section is correspondingly reduced in size and its optimum speed will coincide with that of the axial part. Therefore, the two sections can be combined in one machine with one single rotor running in two bearings only, a design principle well known from the industrial single-shaft gas turbine.

Stage and cooler optimization for the predominant pressure ratios for air between 6 and 8 led to a compact axial-centrifugal compressor with six axial and three centrifugal stages and three pairs of intercoolers. This configuration results in an excellent overall isothermal efficiency due to the higher efficiency of the axial part and the high stage efficiency of the subsequent three wide impellers (see Fig. D-2).


The horizontally split casing consists of six major components ( Fig. D-33). The cast axial inlet (1) and center part (2) cylinder are flanged to the welded centrifugal casing (3). The inlet casing alone or together with the center part can be lifted for the inspection of their internals while leaving the centrifugal casing in its place. The cast blade carrier (7) and diffuser wall (11 ) are also bolted to the centrifugal casing. The discharge volute (4) flanged to the centrifugal casing need not be dismantled when lifting the top half of the latter for the purpose of removing the intercooler tube bundles or inspecting the impellers. For all maintenance operations the external pipe connections remain undisturbed.

The bearings are easily accessible by simply lifting the bearing housing top (6) on the discharge side or the top half of inlet casing (1) and bearing housing (5) on the suction side. The oil and sealing-air connections are located in the bottom half of the respective casings, and any inspection or maintenance work leaves them unaffected. The discharge nozzle forming part of the bottom half discharge volute (4)is normally pointing downwards, but can also be directed horizontally. Two pendulum-type feet and two additional feet attached laterally to the centrifugal cooler casing support the machine on the foundation.

The flow passages to and from the intercoolers follow the same design principle as for the RIK series.


The blade carrier (7) and the short diffuser wall (11) bolted together are flanged to the cooler casing (3) and can expand freely towards the suction side ( Fig. D-33). The double-casing design with outer casing and blade carrier offers various advantages:

++Rigid casing construction; the clearances in the blade duct are not influenced directly by external pipe forces.

++Simple fitting of the blades and assembly of the casing parts; the top half of the casing can be raised without dismantling the blade-adjusting mechanism.

++Possibility of fitting different blade carriers, for adapting the blade channel and thus the compressor characteristics to greatly changed operating conditions.

++Optimal protection of the adjusting mechanism in the space between the casing and blade carrier; the space is kept under suction pressure in order to safeguard the adjusting mechanism against condensation and corrosion attack.

Each of the adjustable stator blades (9)is made out of one piece with a cylindrical shaft. The latter is seated in a bearing bush in the blade carrier ( Figures 11D-37 and 38). The high damping characteristics of this seating arrangement practically excludes the occurrence of dangerous vibration amplitudes associated with the stator blades.

Fig. D-21 Casing Internals Impeller (7), cast iron partition walls (4), vaned diffusors (5). (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-22 RIK skid completely shop-erected, as transported to site as a single-lift package. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-23 Section through an RIK series ISOTHERM compressor (above: vertical section; below: horizontal section)

1 -Casing; 2- Inlet housing; 3- Discharge volute; 4- Partition walls; 5- Diffusors; 6- Shaft; 7- Impellers; 8- Balance piston; 9- Shaft seals; 10- Bearing housing, discharge end; 11- Bearing housing, Intake end; 12- Journal bearings; 13 - Inlet guide vanes; 14 - Vane adjusting mechanism; 15 - Cooler tube bundle; 16 - Water separator; 17- Coupling flange.

(Sulzer Turbo Ltd, Zurich, Switzerland)


The adjusting mechanism is located in the annular space between casing and blade carrier. It’s maintenance-free and does not require any lubrication.

The adjusting mechanism is operated by means of two hydraulic servomotors (10)which are affixed laterally to the bottom half casing. One of the servomotors is equipped with a positioning transmitter and the second operates hydraulically in parallel ( Fig. D-39).

The linear movement of the servomotor piston rods is transmitted directly to the adjusting cylinder (8) by way of two ball and socket joints. The adjusting cylinder of welded design can move in the axial direction and is dry-seated. There is no restriction of heat expansion in any direction. U-shaped guide rings are provided on the inner side in which the adjusting levers are engaged. These rings facilitate the assembly of the adjusting cylinder with the stator blade linkages.

The adjusting levers provided on the end of each stator blade shaft are connected to the guide rings of the adjusting cylinder by means of pivoting sliders. The axial movement of the cylinder is converted into a rotating movement of the stator blades ( Fig. D-40).

The self-lubricating bearing bushes of the blade shafts are seated in the radial holes of the blade carrier. O-ring packings prevent the ingress of contaminants into the stator blade seating.

Fig. D-24 Cross-section through diffuser and return channel. (Sulzer Turbo Ltd, Zurich, Switzerland)

TBL. D.4 Materials of Construction --- Part Material DIN Standard Comparison/ASTM Standard Casing Nodular cast iron*** GGG-40*** A 395*** Compressor inlet Nodular cast iron GGG-40 A 536 Bearing housing Nodular cast iron GGG-40 A 536 Partition walls Cast Iron GG-20 A 48, Class 30 Diffusers: Discs Nodular cast iron GGG-40/1693 A 536 Blades Carbon steel plates HI/17155 A 515, Grade 55 Shaft Low-alloy steel 28 NiCrMoV 8 5 A 470 Impellers Forged steel * * Dummy piston Alloy steel 34 CrNiMo 6 AlSi 4340 Journal bearings Steel with white metal CK 15 + WM AlSi 1015 + WM Inlet guide vanes Stainless steel X 20 Cr 13 Cooler tubes Copper SF-CuF 20 CuNi 10Fe alternatively:** Aluminum brass CuZn 20 A1 B 111/687 F34/1785 Copper nickel alloy CuNi 10 F 29 B 111 C 70600 Copper nickel alloy CuNi 30 F 36 B 111 C 71500 Fins** Copper For all tube alternatives Tube plates top Carbon steel plate HI/17155 A 515, Grade 55 alternatively: Muntz metal CuZn 38 SNAL B 171 C 36500 Tube plates bottom Muntz metal CuZn 38 SNAL B 171 C 36500 Water separators Stainless steel X 5 CrNi 18 9 A 167, Grade 304 Water boxes Cast iron GG-20 A 48, Class 30.


By means of double compressed-air sealing on the suction side combined with a double-walled bearing housing (5) which vents to the atmosphere, preventing any sacking in of oil mist in the event of subatmospheric pressure at the machine inlet ( Fig. D-41 ).


A separate high-pressure control oil unit ( Fig. D-42) actuates the hydraulic servomotor of the adjustable axial stator blades. This control oil unit comprises an oil tank, two motor-driven pumps, a changeover-type twin oil filter, two bubble accumulators, a regulating valve for constant pressure and the necessary instrumentation. All components are mounted on a bedplate and piped up accordingly. In case of failure of the control oil pumps the two accumulators will supply enough oil for a quick and safe reaction of the control elements.

Fig. D-27 Inlet guide vanes with pneumatic actuator.

(Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-25 Bearing housing suction side with special sealing system. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-28 Inlet housing with inlet guide vane linkage system. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-26 Rotor. (Sulzer Turbo Ltd, Zurich, Switzerland)


The basic design principles with their distinct advantages are shown on page 6, Figure 7. Both axial and centrifugal part are forming a single forged monobloc shaft (see Fig. D-44). The axial blades have rhomboidal fir-tree roots and are firmly braced in an exactly defined position in peripheral grooves of the shaft ( Fig. D-38). The three centrifugal impellers of a high flow coefficient and therefore inherent high efficiency ( Fig. D-2) are of the fully welded construction with three-dimensional blades ( Fig. D-3). They are balanced at low speed and subjected to an overspeed test. The assembled and bladed rotor is then dynamically balanced over the whole speed range up to full speed. The sense of rotation of the rotor is clockwise, seen from the suction side.

FIG. D-29 For Inspection of the Intercoolers, each tube bundle can easily be withdrawn Individually. (Sulzer Turbo Ltd)

Design Features


The three pairs of intercooler tube bundles are mounted in a vertical position on each side of the centrifugal casing and bolted to the lower water box. They can freely expand upwards. The upper water box is fixed to the upper tube plate and guided in the top water chamber cover ( Fig. D-43). Round finned tubes ensure excellent heat transfer on the air side. The gap between tube bundle and casing at the exit of the cooler is sealed with a rubber membrane to avoid bypassing of uncooled air. The cooling-water connection and condensate drains are located at the bottom.

FIG. D-30 For mere Inspection of rotor and casing Internals, the Intercoolers need not necessarily be withdrawn. (Sulzer Turbo Ltd)

FIG. D-31 Compressor starting torque with closed Inlet guide vanes (Sulzer Turbo Ltd, Zurich, Switzerland)


TBL. D.5 Technical Data and Dimensions

Technical data, dimensions and weights (dimensions in ram, weights G in metric tons)

E = Gear center distance (average)

H1 = Crane height to lift casing over rotor H2 - Crane height to lift cooler tube bundles over casing H3 -- Crane height to lift casing top half over cooler tube bundles G1 = Casing top half G2 = Heaviest single cooler tube bundle G3 = Bare compressor Ga = Complete skid, max.

mr 2 = Compressor rotor mass moment of inertia in kg m 2, referred to compressor speed, max.

Q = Cooling-water rate of compressor intercoolers in m3/h F = Maximum oil filling of oil tank in liters Main components 1 Compressor 2 Base frame with integrated lube oil system 3 Main oil pump 4 Auxiliary oil pump 5 Twin oil filter 6 Oil cooler 7 Gear 8 Motor 9 Instrument rack l0 Oil mist fan Pipe connections 11 Compressor suction 12 Compressor discharge 13 Cooling-water inlet 14 Cooling-water outlet 15 Compressor condensate drain


1 Casing 2 Inlet housing 3 Discharge volute 4 Partition walls 5 Diffusers 6 Shaft 7 Impellers 8 Balance piston 9 Seals 10 Bearing housing, discharge end 11 Beating housing, intake end 12 Journal beatings 13 Cooler tube bundle 14 Coupling flange

Five frame sizes, providing a complete coverage over the wide flow range from 12 000 to 90 000 m^3/h.

Fig. D-32 Centrifugal casing Internals. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-33 Section through an ARI series ISOTHERM compressor (above: vertical section; below; horizontal section)

1- Inlet casing of axial section; 2- center part of axial section; 3- Casing of radial section; 4- Discharge volutel 5- Double-walled bearing housing, suction side; 6- Bearing housing, discharge side; 7- Blade carrier; 8- Blade adjusting cylinder; 9- Adjustable stator blades; 10- Servomotor; 11 - Short-diffuser wall; 12 - Bladed diffusers; 13 - Partition walls; 14 - Cooler bundles; 15 - Water separator; 16 - Water chamber covers; 17 - Shaft; 18 - Rotor blades; 19- Impellers; 20- Journal bearings; 21- Position of thrust bearing, if fitted; 22- Balance piston; 23- Shaft seal.

(Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-34 Functional and compact design with easy access to all vital parts characterize the ARI series ISOTHERM compressors. (Sulzer Turbo Ltd, Zurich, Switzerland)

TBL. D.6 Materials of Construction

Fig. D-35 Flow passage between axial outlet and entry into the adjacent pair of intercoolers. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-36 Flow passage from Impeller outlet to the adjacent pair of intercoolers. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-37 Adjustable stator blade, rotor blade and fixed stator blade with Intermediate piece.

(Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-38 Fixation of adjustable stator blades. (Sulzer Turbo Ltd, Zurich, Switzerland)

FIG. D-39 Hydraulic servomotor. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-40 Stator blade adjusting mechanism.

(Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-41 Double compressed-air sealing system on the suction side bearing housing.

(Sulzer Turbo Ltd)

Fig. D-42 Power oil unit. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-43 Two-pass version of the intercooler. Finned tubes and highly effective water separators contribute to the excellent efficiency of the compressor. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-44 The single mono-bloc rotor running in only two journal bearings ensures high rotor stability and low vibration level. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-45 Inspection of the internal parts Is made easy by simply lifting the centrifugal casing top half without disturbing coolers and adjacent casing parts. (Sulzer Turbo Ltd, Zurich, Switzerland)

Technical Data and Dimensions (ARI)

TBL. D.7

Fig. D-46 Dimension drawing. (Sulzer Turbo Lid, Zurich, Switzerland)

Fig. D-47 Compressor starting torque with closed stator blades (Sulzer Turbo Ltd, Zurich, Switzerland)

Typical Plant Layout

Fig. D-48 Typical plant layout. Main components 1- Compressor; 2- Gear; 3- Synchronous motor; 4-Exciter; 5- Weather hood; 6- Roll-o-matic filter (1st stage) supply unit; 7- Bag filter (2nd stage); 8- Bypass doors; 9- Suction silencer; 10 - Rubber expansion joint; 11 - Mechanically driven main oil pump; 12 - Motor rotor withdrawal; 13 - Motor cooler; 14 - Discharge silencer; 15 - Nonreturn valve; 16 - Discharge Isolating valve; 17 - Discharge air header; 18 - Blow-off valve; 19 - Blow-off silencer; 20 - Lube oil; 21 - Power oil supply unit. (Sulzer Turbo Ltd, Zurich, Switzerland)

Control system and oil supply


Two main requirements are to be met by the control system of turbo-compressors:

++Compressor safety - to prevent the compressor from operating in an unstable range or at other hazardous conditions

++Process requirements -to adjust the compressor to the demands of the process

The engineering and supply of complete compressor control and safety systems ensures the optimum protection of the compressor and the plant.

Thanks to the strict use of standard signals, the control sys- tem can be integrated into other systems without difficulty.

It allows remote control, automation of starting and stopping and can be linked with distributed control systems (process computers).


Antisurge control

The stable operating range is defined by the characteristic curves C and limited by the surge line S ( Fig. D-49). Operation under surge conditions, occurring on the left side of this surge line, is avoided by an antisurge control system. It measures flow and pressure and is designed to closely follow the actually measured surge line with a predetermined safety margin.

As soon as the operating point approaches the response line (L), the controller progressively opens the antisurge valve as a function of the difference between the minimum stable flow of the compressor at a given pressure and the flow required by the process. This valve is either a blow-off valve (air) or a bypass valve (nitrogen, oxygen) followed by a bypass cooler.

As a protecting device the system has to act independently of any other control system and must not be used for pressure or flow control. Limitation of temperature, pressure, axial displacement Under certain circumstances, external influences or other irregularities may lead to undesired changes of the normal level of gas and bearing temperatures, pressures, axial displacement of the rotor, etc. A reliable interlock, alarm and shutdown system must protect the compressor train from possible damage under such conditions.

Auxiliary component control Auxiliary component control assures a safe supply of lube and control oil.


Suction pressure

Constant suction pressure to adapt the compressor flow to an upstream production unit.

Discharge pressure Constant discharge pressure in cases where chemical reactions or physical processes have to take place at a clearly defined pressure, or where the compressor flow has to be adapted to a fluctuating down-stream demand.

Flow Constant mass flow control corresponding to a constant plant output. The above-mentioned process control systems can also be combined via selecting relays, e.g. flow control with discharge pressure limitation.


The typical P & I diagram shows all the necessary control and monitoring equipment for the safe operation of an isotherm compressor. Additional control and monitoring equipment can be added on request.


For the RIK series the lube oil system is normally integrated in the compressor skid and comprises the same components as the Sulzer standard oil supply unit shown in Fig. D-50.

The separate standardized lube oil units for the RIO and ARI series comprise the oil tank, auxiliary oil pump, twin oil filter, selectively single or twin oil cooler, oil mist fan, oil filling sieve, electric heater, automatic start-up testing device, oil pressure control valve, and all the valves, apparatus and instruments necessary for the safe operation of the compressor set ( Fig. D-50). The main oil pump is normally mechanically driven by and attached to the speed increasing gear or the steam turbine. The lube oil system is designed to supply the lube oil for compressor gear and driver.

Fig. D-49 Characteristics of a turbo-compressor p - Discharge pressure; V - Flow rate; C - Compressor characteristic curves; S - Surge line; L - Limit flow.

(Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-50 Standard lube oil supply unit. (Sulzer Turbo Ltd, Zurich, Switzerland)

Typical Compressor Plant P &l Diagram --


FIG. D-51 Air compressor, type RIK 56, at the mine of the Gold Fields of South Africa Ltd., Johannes- burg, compressing 51 000 m3/h from 0.825 to 9.5 bar. Power Input 4000 kW. (Sulzer Turbo Ltd, Zurich, Switzerland)

FIG. D-52 Air compressor, type RIK 71, Installed In one of AGA's Air separation plants In the USA. 65 000 Nm3/h are compressed from 0.97 to 6.76 bar; power Input 5535 kW. (Sulzer Turbo Ltd)

Fig. D-53 Nitrogen compressor train consisting of an Isotherm compressor, type RIK 56, with RZ 35-6 booster, Installed at Union Carbide's air separation plant Prentiss, Canada. The 7900-kW electric motor driven set compresses 37 420 Nm2/h of nitrogen from 0.91 to 47.1 bar. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-54 Compact Isotherm process air compressor, type RIK 63, driven by a tall gas expander E 40 and a steam turbine with a combined power output of 8200 kW. The set compresses 50 000 Nm3/h of air to 10 bar in a South Korean nitric acid plant. (Sulzer Turbo Ltd, Zurich, Switzerland)

FIG. D-55 Oxygen plant Burns Harbor, USA, of Union Carbide’s Linde Division, each Air separation line supplying 2000 st/d of oxygen to neighboring steel works. The two motor-driven Air compressors, type ARI 80, deliver 300000 Nma/h each at 7.2 bar, absorbing 25.2 MW. The two oxygen compressors supply 55 000 Nm3/h each at 34.2 bar; power Input 9050 kW. The main Items of this compressor plant are: 1 - Oxygen compressor I; 2 -Air compressor I; 3 -Air filter house I; 4 - Air aftercooler for Air compressor I; 5- Oxygen compressor II; 6- Control room II; 7- Air aftercooler for Air compressor II; 8 - Electric motor II; 9 - Noise attenuating enclosure for gearbox; 10 -Air compressor II; 11 - O11 reservoir II; 12 - Air filter house II. (Sulzer Turbo Ltd)

Fig. D-56 These two air compressors, type ARI 56, Installed ad Anglo American's Vaal Reefs mine, South Africa, are the biggest mining compressors of the world. Each machine supplies 170 000 m3/h at 9 bar and is driven by a 15-MW synchronous motor. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-57 Skid-mounted air compressor unit, type RIK 56, with R 35-5 and steam turbine arriving at an ammonia plant in Cartagena ( Spain). Compressor capacity 34 900 Nm3/h, discharge pressure 34 bar. (Sulzer Turbo Ltd)

Fig. D-58 One of two Isotherm compressors, type ARI 63, Installed In the steelworks of Diefflen/Dillingen, FRG, shown during erection on site. The compact units with Integrated coolers allow their Installation also at locations where space is very limited, still ensuring easy access to all vital parts for maintenance. Each compressor delivers 180 000 Nm3/h of atmospheric air at 7.1 bar. Power Input 15.4 MW. (Sulzer Turbo Ltd, Zurich, Switzerland)

Fig. D-59 This motor-driven. 12.6-MW compressor, type ARI 63, delivers 139420 Nm3/h of air at 7.17bar to Airco-Cryoplants' air separation installation supplying oxygen for the Coolwater coal gasification plant at Daggett, Cal., USA. The eyngaa produced serves as fuel for a 100-MW cogeneration power plant. (Sulzer Turbo Ltd, Zurich, Switzerland)




Gas Seal Design

FIG. E-1 (Dresser-Rand, Olean, NY) FIG. E-2 (Dresser-Rand, Olean, NY)

Gas seal designs are a critical feature in compressor design, operation and maintainability. Manufacturers constantly seek improvements in this area as is indicated by the following extracts from a technical paper.

The purpose of this Sub-section is to emphasize two of the main technical features incorporated in the design of the D-R Gas Seal:

++The groove pattern (generates high film stiffness and optimum pressure distribution)

++The "L" Sleeve design (reduces the risk of "hang up" and therefore cuts operational costs). Gas Seal Principle Fig. E-1 shows a simplified cross-section drawing of a D-R gas seal where the main active parts are:

++The rotating seat, also called tungsten carbide ring

++The stationary seat, also called the carbon ring

++The pusher sleeve, also called "L" sleeve Figure 1 I E-2 shows, in parallel, the essential parts of a D-R gas seal, and the detailed representation of the closing and opening forces which determine the equilibrium of the seal faces.

When calculating the balance of forces on the carbon face for a given gap, we take into account:

++the opening forces (OF): sum of (pressures inside the gap x surface)

++and compare that to the closing forces (CF): sum of (pressures at the back of the carbon face x surface on which they apply + spring force (Fs) -friction force (FI") If the gap increases, then the opening forces will decrease; if the gap decreases, then the opening forces will increase. The film of gas acts just like compression springs: If the speed increases, then the opening forces will increase as the grooves will generate more lift.

Conversely, we can also adjust the closing forces. If we increase the hydraulic diameter, then the closing force will diminish; and if we lower the hydraulic diameter, the closing forces will increase. This is called 'playing with the balance ratio'.

This can be put in simplified equations as follows CF = Pin • S l + Pout X 52 "+" Fs OF = k (rpm, press, groove geometry, temperature, gap, gas characteristics)

In the steady-state situation OF--CF. Thus the solution of these equations is to calculate the gap (operating gap) which exactly balances the opening forces and the closing forces Groove Pattern and Pressure Profile

Considering the above equations, while it’s simple to calculate the closing forces, the calculation of the opening forces is a complex function of several parameters, which are not independent. Therefore Dresser-Rand have developed a computer code that will calculate pressure, speeds and temperature distributions at every point within the interface. This program iterates until the operating parameters are established and stable; in fact it finds the gap at which the opening forces exactly balance the closing forces. Of course various groove geometries can be analyzed and then optimized in order to provide the highest reliability and performance to the gas seal.

Ideally, a gas seal would like to have a minimum gap in order to minimize the leakage rate (the gas seal leakage is approx, a function of the gap raised to the power 3). However, a smaller gap gives a higher risk of accidental contact between the two faces of the seal. A small gap as well as the highest possible gas film stiffness will be the optimum. A uniform pressure distribution between the faces is important, since it reduces the local deformation of the parts. Therefore, let's discuss pressure profiles, balance of forces, and gas film stiffness.

It’s clear that the layer of gas trapped between the rotating seat and the carbon face of the gas seal changes pressure as it goes from the OD (at seal supply pressure to the ID (at flare pressure or atmospheric pressure). The grooves in the rotating seat alters the normal pressure decay and generates zones of overpressure.

Using the pressure at each point of the interface, calculated with the above mentioned calculation code, it’s possible to generate a 3D representation of the pressure profile. One example is shown in Fig. E-3.

From the same computation output it’s also possible to show the isobar curves ( Fig. E-4) which highlight the over-pressurized areas responsible for the separation of the seal faces.

Comparing the various pressure profiles and isobar curves obtained with different groove patterns, it’s possible to select a groove shape which provides high pressure rise and as uniform as possible pressure distribution. The patented groove pattern used by Dresser-Rand satisfies these two important criteria.

As an example, Figures 11E-5 and 6 show isobar curves at different operating conditions (high speed/low speed and high pressure/low pressure)

FIG. E-3 (Dresser-Rand, Olean, NY)

FIG. E-4 (Dresser-Rand, Olean, NY)

FIG. E-5 (Dresser-Rand, Olean, NY)

Gas Film Stiffness

Gas film stiffness is a major parameter for gas seals as it can be used to evaluate the ability of a gas seal to resist sudden positional changes (surge for instance), or also to compare two seals with different groove geometry Let's define what the gas film stiffness is: At a given operating gap an opening force (OF1) is generated within the interface, if the gap is forced to close (or open) by say 1/100 of its value (gap/100), a new opening force will develop (OF2).

In a similar way, springs are calculated with the formula:

dF = S x dX, hence S = dF/dX

in this case S = stiffness; dF = (OF2- OFI); and dX = gap/100.

The gas film stiffness is (OF2- OF1) x 100/gap.

As an order of magnitude, stiffness of more than 3 kN/micron for a medium size of seal (4.875 inch diam) at 50 Bar and 11500 rpm are typical. See Fig. E-7.

In imperial units, this would be about 175 x 106 lbs/in (at 725 psi). This is a very high value, but necessary to avoid any contact between the two faces which at the same time are separated by about 4 microns (0.16 thousandth of an inch). At higher pressure the stiffness values are even higher, which is necessary as the gap will also be smaller. In fact, the stiffness value gives an indication of how the seal can withstand axial forces (especially abnormal forces due to vibrations or upset conditions.)

Incidentally, it’s possible to use the same program to calculate the behavior of seals having a symmetrical groove pattern (bi-directional). The results demonstrate a significant decrease of gas film stiffness together with a smaller gap. This is why bi-directional seals are not as forgiving as uni-directional seals.

The Hang-up Syndrome

Considering again the seal equilibrium, let us analyze now the event of a very low sealing pressure (for instance start up conditions before pressurizing the compressor). See Figure llE-8.

FIG. E-6 (Dresser-Rand, Olean, NY)

When the seal is depressurized, only the spring force can close the seal gap.

In this situation, if the shaft of the compressor has a small axial displacement (for instance differential thermal expansion between shaft and compressor casing) the friction between the O-ring and the seal housing or the deformation of the O-ring itself, may prevent the spring closing the seal. A much larger gap (50 to 100 times the normal gap) may then appear between the two faces of the seal. The leakage through this interface could then be so high that pressurizing the unit becomes impossible.

This is known as the 'hang-up syndrome'. The only remedy is to disassemble the compressor end, remove the seal cartridge, fix the gas seal and reinstall it. This, of course involves unnecessary down time and high maintenance costs, usually at a time when there is an urgent need to have the compressor up and running.

Dresser-Rand has developed and incorporated in the D-R Gas Seal, an "L" sleeve design (patent pending) which drastically reduces the risk of "hang up".

FIG. E-7 (Dresser-Rand, Olean, NY) FIG. E-8 (Dresser-Rand, Olean, NY)

The 'L' Sleeve Design

As explained above, the hang-up situation is caused both, by the tendency of the O-ring to stick or to extrude, and by the simultaneous lack of gas pressure, pressure which in normal operating conditions, is the major contribution to the closing force.

The Dresser-Rand "L" sleeve design addresses these two causes: (a) Reduced risk for extrusion The balance diameter O-ring is located remote from the hot area of the seal, thus it’s subject to a lower temperature. The clearance may be better adjusted (smaller) since the housing and the sleeve are of the same material (same thermal expansion coefficient and therefore less risk or tendency for extrusion). (b) Increased closing force due to the gas pressure (small but existing).

The installation of the O-ring in the sleeve (as opposed to its installation in the housing) tends to decrease the balance diameter, thus slightly improves the closing forces due to the gas pressure (see Fig. E-9).

FIG. E-9 (Dresser-Rand, Olean, NY)

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