Some compressors in certain applications have one or more incoming side streams (or side-loads) that introduce side-stream flow to mix with the core flow. In other words, a radial side-stream inlet is utilized in multi-impeller centrifugal compressors to introduce additional gas into the mid-stage of a compressor. This design enables optimization in some process plants. However, the flow distribution after the junction of the side-stream and the main return channel of the upstream section can significantly affect the performance of the next section of compressor. In most cases, the pressure levels at these side streams should be accurately predicted to meet the theoretical performance.

The design and operation of compressors with side streams has always been difficult because flow entering the compressor needs to be mixed with the core flow already in the compressor (compressed in section-1) in a manner that does not degrade the aerodynamic performance of surrounding sections. Minimizing losses and ensuring proper mixing of the side-stream flow and core-flow are required to ensure effective performance of a compressor of this sort.

Key considerations for an application involving compressors with side-stream flow include:

  • The plant process typically dictates the side-stream flange pressure.
  • The impellers upstream of the side-stream should achieve the necessary pressure for the core-flow.
  • The side-stream flow is typically at a temperature different from compressor core-flow temperature.
  • Process plants need some degree of operational flexibility, i.e., the ability to support variations in core-flow conditions and side-stream conditions.

Modern process plants are usually specifying more than one operating conditions (sometimes two or more mandatory conditions), which can present challenges to compressors with side streams. Adding to these issues, modern compressors should operate with much higher impeller tip speeds and inlet Mach numbers compared to previous generations of compressor technology.

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Likewise, side-streams are now commonly used with medium-sized and large-sized compressors (e.g., 1–60 MW). The costs associated with corrections on side-stream compressor applications — any repair, inefficient operation and power losses — would be very high. Further, compressors with side-streams are usually sensitive and difficult machines to operate. In particular, performance curves (performance, pressures, efficiency, etc.) are sensitive to suction inlet conditions, side-stream inlet conditions, volume flow ratio (side-stream volume to core-flow volume), and side-stream losses.

Compressors with side-stream often perform far below expectations. The main reason is that side-stream compressors are different than conventional compressors (without a side stream) in design and operation. For a compressor with side-streams, tighter tolerances in operating conditions must be identified and applied. Uncontrolled variations of suction pressure or side-stream pressure are major problems. Another major issue is the variations of the flow ratio (side-stream flow to core flow).

Compressor side-stream & mixing section

In many centrifugal compressors, the portion of the side-stream system from the flange connection to the mixing section is similar to a compressor inlet. A great challenge is ensuring the mixing section provides circumferentially uniform merged-flow with minimum losses. This can be achieved via effective design of the inlet configuration and mixing section of the side-stream compressor, such that the impeller can be maintained over the whole operating range of the compressor. In many cases, while there are mixing sections provided for the machine, the flows do not completely mix before entering the impeller of the section-2 (the impeller after the mixing section). Non-uniformity at inlet flow to the impeller is a key concern for compressor performance.

Another consideration is a compact design of the side-stream mixing section. Specifically, a mixing section with less axial length is always preferred. This axial length can affect the compressor rotor dynamics, the rotor length, and overall compressor performance (design, cost, operation and reliability).

The side-stream exit pressure (side-stream pressure at inlet of the mixing section) is mainly a function of the side-stream flow velocity and the pressure in the previous section of the compressor. Conversely, the side-stream flange pressure can depend on side-stream flow velocity and section-1 exit conditions. Therefore, the side-stream inlet flange volume flow is a function of the flange pressure, compressibility (at that pressure and temperature), and mass flow. Some engineers assume that the side-stream flange pressure is equal to the inlet pressure of the first impeller of the section-2. This is not true. The mixed total pressure can be estimated on a mass-averaged basis using side-stream total pressure and return channel exit (section-1) total pressure.

In some designs, because of the lower static pressure at the side-stream exit, the total pressure at the return channel exit is lower even though the velocity upstream is roughly matched for both streams. The pressure usually increased at the returned channel, but it can be reduced again at the side-stream exit. While simple in nature, this can have an impact on the prediction of pressure at the side-stream exit and overall mixing section performance. Based on basic flow principles, it was usually assumed that the static pressure was equal between both flow streams at the mixing location. While this assumption might be used in rough calculations, the accurate simulations have shown this assumption is not true. In fact, there are some points (for instance, the shroud wall at the exit of the return channel and the hub wall at the exit of the side-stream) where this assumption is true. However, this is not valid for all locations. There are pressure changes that occur from a side-stream flange to the inlet of the next section (section-2), which depend on different factors. The first factor is the frictional losses due to side-stream geometry, from side-stream flange, to plenum, mixing section, etc. The other factors are pressure changes due to local curvature and geometry details of mixing section.

Another consideration is the operation of a compressor with side-stream at part-load. The performance of the side-stream path, the mixing section, and the downstream impellers when machine is operated at part-load or even off-design operating flows will result in additional compressor performance problems. In such a case, both velocity profiles and pressure variations in various parts (side-stream path, mixing section, etc.) will be different, resulting in complicated flow patterns and complex aerodynamics behaviors.

Side-stream compressor sectional performance

Compressor with Side-Stream SIEMENS
Figure 1. An example of a centrifugal compressor with side-stream.

Typically, vendor-supplied compressor performance curves reflect flange-to-flange performance because that is what the plant engineers need to evaluate the proper operation of their system (for example, from suction to side-stream and from side-stream to discharge). In many cases, compressor vendors also provide the overall compressor performance curves and performance curves for section-1 (from suction to side-stream) and section-2 (from side-stream to discharge). Flange-to-flange data, if not interpreted properly, can lead to false conclusions about the relative performance of individual sections of a compression system. For example, if side-stream losses from the flange to mixing section are attributed to the downstream section (section-2), it will cause the downstream section (section-2) to appear low in performance, while the upstream section (section-1) will show high performance levels.

An important behavior to consider is when the actual side-stream pressure is different than the rated side-stream pressure. When the side-stream pressure is slightly higher than the outlet pressure of the section-1, it (section-1) shows a relatively higher performance compared to the real section-1 performance. In some cases, the performance of section-1 could be considerably higher than section-2. In other words, when performance is determined flange to flange, the higher flange pressure is seen as extra “pseudo” work done by the section (since this is an inconsistent thermodynamic volume), resulting in a relatively high sectional efficiency (section-1 shows a relatively high efficiency). This is the case for many compressors with side-stream because the whole system is designed (or specifically the side-stream location is selected) in a way aimed at maintaining the real side-stream pressure (side-stream inlet) above the section-1 outlet pressure because if the real side-stream pressure becomes lower than the section-1 outlet pressure, the flow may (theoretically) reverse.

This is particularly important when the side-stream mass flow is much lower than the mass flow of core-flow. The above-mentioned reasons explain why in many simulations, calculations, and performance reports related to compressors with side-stream, the section-1 shows a relatively good efficiency and section-2 appears low in efficiency.

When the side-stream pressure is slightly lower than the outlet pressure of the section-1, the section-1 might show a relatively lower performance compared to section-2. This could occur in some recycle services where the side-stream is actually a recycle flow with a mass flow around two to five times the core-flow. The side-stream (recycle) flow is returned from downstream of the compressor to be slightly compressed and recycled to the downstream. Such a recycled side-stream compressor can present unique challenges for design and operation. This kind of machine usually needs very special operating procedures and very fine adjustments regarding operating pressures and flows.

Side-stream compressor performance & operation

An important consideration is the sensitivity of compressor overall performance curve and sectional performance curves to the flow ratio (ratio of side-stream flow to core-flow) and to the inlet conditions. Variations in the flow and pressure of the main suction and side-stream inlet should be controlled within tight tolerances. For compressors with side-stream, the ASME PTC-10 code stipulates limits on flow ratio (side-stream volume to core-flow volume). The acceptable variation in volume flow ratio in a compressor with a side-stream is +/-5 percent as per ASME PTC code. The requested side-stream pressure tolerance is not usually specified in compressor codes, but it should be around +/- 2 percent (or sometimes +/- 2.5 percent).

Variance in flow ratio impacts the velocity levels where the two streams merge. Significant variation in the velocity profile upstream of the impeller changes the incidence on the blade leading edge of the following impeller. This change in incidence leads to a change in sectional performance, overall compressor performance, and train efficiency.

If the flow ratio is varied between 95 percent and 105 percent (tolerances specified in the ASME PTC-10 code), the flange pressure changes accordingly, resulting in a change in the sectional performance and overall performance. The sectional efficiency varies sometimes more than +/- 4 percent. This variance reduces as the flow is decreased toward surge. However, as the flow is increased toward overload, the variance increases up to higher values (even in some cases three to five times compared to variations near the surge zone). For example, in a case study for a compressor with side-streams, a +/- 5 percent variation in flow ratio and associated changes (pressures, losses, etc.) the sectional efficiency reduced by approximately 3.4 percent at the rated flow; at allowable operating points at the end of curve (near the overload, i.e., choke) efficiency reduced by approximately 5.5 percent. These results show that the variation in sectional performance and overall performance due to flow ratio is highest at operating points near the overload and lowest at operating points near the surge. For compressors, which work most of the time at high flow (flow higher than rated flow and discharge pressure lower than rated pressure), the effects of flow ratio deviations (and subsequently efficiency reduction) would be higher than compressors working on the left-hand side of curve (say with sufficient margins near the surge).

In the compressor codes (such as API 617), the compressor power variations should be maintained below 4 percent. However, in some applications, clients require tighter tolerances on the consumed power, sometimes below 2 percent (or even below 1.5 percent). Some studies suggest while +/- 5 percent of volume flow ratio is suitable for performance test as per ASME PTC-10 code, some plants require tighter control on the consumed power and compressor performance (lower limits on the efficiency), the tolerances on the volume flow ratio could be tighter than ASME PTC code. In this regard, +/- 4 percent tolerance can be suggested on the volume flow ratio (ratio of side-stream flow to core-flow) of compressors with side-streams.

Based on operational experiences, acceptable variations in volume flow ratio and inlet pressures (suction inlet and side-stream inlet) of a compressor with side-streams are +/- 4 percent and +/- 2 percent, respectively.

Another important consideration is the required adjustments for the flow ratio (side-stream flow to core-flow). Compressors with side-stream usually require fixed side-stream flange pressures. For a compressor with some variations in operation conditions (even small changes), it is not usually possible to achieve optimum operation by keeping the flow ratio (side-stream to core-flow) constant. The flow ratio has to be adjusted according to the inlet flow at which the machine is operating in order to get the required corresponding side-stream flange pressure.

Thermal changes have also been known to cause some instability in centrifugal compressors. An example is the thermal-induced change of residual unbalance. For compressors with side-streams, which feature temperature differences between core-flow and side-stream flow, the thermal behavior of the compressor is usually more complex and requires more attention.

Heavy gas challenge

Heavy gases (such as heavy hydrocarbon gases, e.g., propane, propylene, MTBE (Methyl Tertiary Butyl Ether)) have very low gas sonic velocities that produce high Mach numbers in the aerodynamic flow paths. For these services, because of high Mach number, high flow coefficient stages have very narrow flow maps characterized by limited surge and choke margins. Compressors with side-streams in these services present great challenges. This is to say, the side-streams and associated mixing further complicate design, operation and performance prediction because the pressure, temperature and flow conditions at each one of these side-streams, as well as at the exit of the machine, require stringent tolerances to optimize the overall efficiency and performance. As such, it is important to accurately model, design, operate, and monitor performance characteristics of any impellers, diffusers, return channels, and generally all flow-path components in a compressor with side-stream.


Amin Almasi is a senior rotating machine consultant in Australia. He is a chartered professional engineer of Engineers Australia (MIEAust CPEng – Mechanical) and IMechE (CEng MIMechE), holds bachelor’s and master’s degrees in Mechanical Engineering, and is a registered professional engineer in Queensland. He specializes in rotating machines, including centrifugal, screw, and reciprocating compressors, gas turbines, steam turbines, engines, pumps, subsea, offshore rotating machines, LNG units, condition monitoring, and reliability. Almasi is an active member of Engineers Australia, IMechE, ASME and SPE. He can be reached at