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MATCHING THE CONFIGURATION TO THE APPLICATION
Back-to-back compressor. Courtesy of Dresser-Rand, a Siemens business
By Mark Sandberg
Complex compression trains may be supplied in series and parallel arrangements to achieve the total flow and head requirements of a specific application. The most fundamental configuration of a centrifugal compressor is the straight-through design.
This layout is composed of one or more impellers, aligned in the same direction, contained within a single casing fitted with a single inlet nozzle and a single discharge nozzle to accommodate the gas flow. The number of stages contained within the section is a function of the produced head requirements. However, there are practical limitations to the number of stages that can be included.
One of these limitations is connected with the resulting discharge temperature that is a function of the overall pressure ratio, compression efficiency, and thermophysical properties of gas. This temperature limitation may be due to material temperature limits of components within the compressor, or gas temperature limits imposed by other equipment or the process in which the compressor is operating.
An important limitation in the number of stages allowed in a single, straight-through configuration is associated with lateral rotordynamic considerations. Current beam-style designs restrict the number of stages in a single casing to 10 or less. However, this may also be impacted by the magnitude of the flow coefficients of the individual stages.
Beam-style compressors are configured with all of the impellers and a balance piston, if applicable, on the rotor located between two radial bearings. Lateral rotordynamic stability is one of the primary considerations in limiting the number of stages to 10 or less.
Regardless of the number of stages provided in a design, it is good practice to provide a preliminary stability screening using the evaluation method of the critical speed ratio (CSR) plot provided in API 684, “API Standard Paragraphs Rotordynamic Tutorial: Lateral Critical Speeds, Unbalance Response, Stability, Train Torsionals, and Rotor Balancing.”
This analysis examines both the rotor flexibility (in terms of the rotor operating speed to first rigid critical speed ratio) and the average density of the gas which impacts the magnitude of the aerodynamic excitation forces generated. Other similar figures generated through the years have replaced average gas density with the product of discharge pressure and pressure differential across the compressor.
The information required to evaluate rotordynamic stability using this plot is normally available on data sheets for API 617, “Axial and Centrifugal Compressors and Expander Compressors for Petroleum, Chemical and Gas Industry Services.”
Another useful parameter that can be calculated using data sheet information is the bearing span-to-impeller bore diameter ratio of the rotor. It is more likely that lateral stability is provided if this ratio is 10 or less. Values above 10 should be evaluated in more detail, and any value above 12 should be considered for an API 684 Level II stability analysis.
The pressure profile that is present around each impeller results in an unbalanced axial thrust force in the direction from the back disk towards the eye of the impeller. A summation of these unbalanced forces represented by each impeller is balanced by the combination of a balance piston at one end of the rotor and a thrust bearing.
The balance piston is provided with a cross-sectional area where discharge pressure is imposed on one end of the piston and suction pressure on the other, partially negating the unbalanced axial force due to the impellers.
Variants of the beam-style, single-section, straight-through design deserve identification. One of these is the overhung, single-stage design that is commonly used in pipeline applications where a relatively low-pressure ratio is required.
The layout of this design includes a single impeller placed outside of the bearing span. The number of impellers included is normally limited to a single impeller due to overhung moment influences on lateral rotordynamic behavior.
Another variant is the integrally geared compressor. In its simplest form, a single impeller is directly connected to the end of a pinion which is driven by a bull gear to a specified speed. These impellers are often of an open design.
The vanes are connected to the back disk, but there is no cover provided that encloses the flow path on top of the vanes. Although the lack of an impeller cover increases leakage, this is offset by the increased head generating capability due to higher allowable tip speeds when compared to a shrouded impeller.
More complex versions of the integrally geared compressor exist: Impellers are attached to both ends of the pinion and multiple pinions are associated with a single bull gear. The rotational speed of each of the pinions can be optimized to enhance the overall efficiency of the compressor.
There are a substantial number of applications where intercooling of the gas is required to maintain temperatures at acceptable levels. One solution is to provide multiple, single-section casings with inter cooling provided between the casings. An alternative is to provide multiple sections within a single physical casing. This is known as a compound section configuration.
One benefit is reduced footprint, since a single casing requires less space than two casings with double the amount of radial bearings, as well as the space required by a coupling between two separate casings. Additionally, process requirements may dictate the need for a side stream addition or extraction to be available at some intermediate pressure of the overall compression application.
But there are some drawbacks to this design. Inlet and discharge nozzles for each section must be supplied on the casing. Adequate spacing to accommodate these nozzles has to be considered when the equipment supplier lays out the design. Although there are no theoretical limitations on the number of sections contained within a single casing, limits exist on the space available in a given casing.
A pressure differential also exists between the discharge of a given section and the suction of the following section. This is due to pressure losses that exist in the piping and equipment between two sections.
Differential pressure within the compressor casing results in the potential for leakage from the lower-pressure section discharge into the higher-pressure section suction. This results in an overall loss of efficiency. Some type of controlled leakage internal seal must be provided to limit leakage and increase compression efficiency.
The addition of this seal requires space. A reasonable rule of thumb is that each seal displaces a potential stage of compression. Accordingly, a reasonable estimate of the maximum number of stages available in a two section compressor is nine. This reduces to a maximum number of eight for a three-section machine. A compound compressor with more than three sections, then, is only feasible for relatively low compression ratios per section.
One final characteristic of the compound design concerns the transient operation of the machine. Upon shutdown, a compressor will settle out to an equilibrium pressure somewhere between the suction and discharge pressures. This is dependent on a number of factors, primarily relative suction and discharge volumes associated with the compressor and its connected systems.
This settle-out pressure calculation is straightforward for a single-section casing. For a casing with multiple sections, though, each section will settle out to a different pressure level, then all sections reach a common settle-out pressure due to internal leakage. Associated piping, vessels, and other process equipment must be designed to accommodate this combined settle-out pressure.
A unique version of the compound design is the back-to-back configuration. The stages of the first section are oriented within the casing opposite to the second section. Generally, the eyes of the impellers of each section are oriented towards the shaft ends of the casings. This configuration is similar to the compound design. It can provide intercooling and mass flow addition, or extraction between the sections, as an option.
One advantage of the back-to-back design is its inherent characteristic to reduce, and roughly balance the axial thrust force generated in the stages of each section. Since the two sections are oriented in an opposite direction, unbalanced axial thrust forces act in opposite directions. This is a distinct benefit in high-pressure, high-density compression applications, such as gas injection services where unbalanced thrust forces can be substantial.
In such an application, the duty and size of the thrust bearing could be prohibitive without the balancing feature. A close clearance, controlled leakage seal must be provided between the two sections. Unlike the balance piston required in the other designs, this seal (commonly known as the division wall seal) is only subject to about half of the overall differential pressure of the two-section casing.
This reduced pressure difference can minimize internal leakage. But it can be offset by a possible need for increased clearance. This is due to the location of the seal near the center of the bearing span with the accompanying increased deflection of the shaft in this location.
Further, the location of the division wall seal near the center of the rotor has been a challenge with the potential for aerodynamic excitation. Technological and design developments over recent years have reduced this issue with the introduction of hole pattern seal designs and swirl brakes.
Final settle-out conditions of the back-to-back design are similar to the compound configuration: Each section settles out to a different pressure level followed by an equalization of pressure between the two sections to a common settle-out pressure. It is critical to design piping and equipment to this common settle-out pressure to prevent over-pressurization, particularly on the suction side of the lowest pressure section.
An additional phenomenon observed with the back-to-back design concerns leakage across the division wall seal. It has an impact on the flow leaving the first section and entering the second section. This flow is the sum of the first section flow rate, any side stream flow, and division wall leakage, which may impact inter-stage process piping, process equipment and intercooler duty and design.
Significant inlet volumetric flow rates can result in excessive impeller flow coefficients. Beyond a certain point, it is beneficial to reduce the flow into an individual section. One way to accomplish this is to provide parallel sections in separate casings. An alternative is to select a double suction compressor design.
This design is geometrically similar to the back-to-back configuration, but both sections are of equal design. The inlet flow rate is split in half externally and introduced into the casing through two suction nozzles at each end of the casing. Upon passing through the two equal sections, the flow is combined into a single discharge nozzle.
Given that the two sections of the compressor are oriented in separate directions and are aerodynamically similar, there is theoretically no net axial thrust produced. Since there can be differences in both internal and external suction losses and manufacturing tolerances between the two opposed sections, some small amount of axial thrust is anticipated and a thrust bearing is included, albeit limited in capacity.
The double suction configuration is useful for large volumetric flow rates that are often associated with low inlet pressures. It should be assumed that the maximum number of stages per section is limited to no more than four. This is due to the presence of dual-suction inlets and a combined discharge.
Although possible, this machine design is probably limited to a single section of compression with no more than two suction nozzles and a single discharge nozzle.
The sideload configuration is also similar to the compound design (Figure 7). Multiple sections are oriented in the same direction and contained in a single casing. However, there are no intermediate intercooled flows that leave and re-enter the casing.
One or more sidestream flows may be introduced or extracted from the casing. But all or most of the flow from the preceding section does not leave the casing. Sidestream flows that are introduced into the casing mix with the discharge of the preceding section. Cooling may occur through the mixing of the predominately lower temperature sidestream.
The sideload compressor design is well-suited for refrigeration applications where the refrigerant is introduced at progressively lower temperatures to the lower pressure sections. These sidestreams originate from economizers that operate at intermediate pressure levels in order to increase the overall efficiency of the refrigeration process.
Gas temperature increases in a given section are reduced by the mixing of the sidestream flow into the suction of the following section of the compressor. This maintains gas temperature throughout the machine at reasonable levels.
It is possible to extract flow from one of these sidestreams. But this is an exception. Limitations on the potential number of sections and stages per section exist for the sideload design. Assume as a rule of thumb that a limit of ten stages is allowed in a single casing, and that any sidestream nozzles are roughly equivalent to a stage. Thus, there are limitations to the number of sections that can be contained.
There are further complications associated with sideload compressors, such as process control. Generally, each section of compression is protected from surge through a recycle line which prevents the volumetric flow rate from falling below a prescribed level. This is complicated in the sideload configuration because recycle flow can only be obtained from the final discharge.
The relative effects of the flow capacities of each section and the preceding sections influence the selection of these minimum volumetric flows. It is also more difficult to monitor the aero-thermodynamic performance of these machines since mixture temperatures of the internal gas flows are not measured. This means individual section performance can only be estimated based upon predicted section efficiency.
More centrifugal compressor design configurations exist such as isothermal designs. But they tend to be unique and limited in application. Those described above are most commonly observed. Further, combinations of configurations may be applied to potential applications.
Author: Mark R. Sandberg, P.E. is the owner of Sandberg Turbomachinery Consulting. For more information, read the paper he presented on this
subject at the 2016 Turbomachinery Symposium: Centrifugal Compressor Configuration, Selection and Arrangement: A User’s Perspective, which is available for download at no charge from the Turbomachinery Symposium website at https://oaktrust.library.tamu.edu/handle/1969.1/158832