Optimizing Peak Efficiency and Flow Range in Pipeline Centrifugal Stages

Transcription

1 Optimizing Peak Efficiency and Flow Range in Pipeline Centrifugal Stages Authors: James M. Sorokes, Jason A. Kopko, José L. Gilarranz, Andrew J. Ranz Affiliation: Dresser-Rand North American Operations, Olean, N.Y. USA INTRODUCTION Traditionally, operators of centrifugal compression equipment have placed a high priority on achieving peak efficiency or minimizing horsepower at one specific operating condition. Often called the design point, this condition corresponds to the flow rate and pressure ratio at which the end user expected to operate the compressor most of the time. In response, compressor designers focused much of their energy on developing centrifugal staging that would provide very high efficiencies. However, quite often, the high efficiency was achieved at the expense of overall operating range; i.e., stonewall (overload) to surge / stall. In recent years, greater emphasis has been placed on achieving wider operating range while maintaining superior efficiency levels -- especially true in the pipeline industry where demand varies greatly during the year. In the past, some pipeline stations would require so-called winter and summer bundles (configurations); i.e., the compressor internals would be changed between the winter and summer pumping seasons. This resulted in additional cost for the operator and introduced the risk of damage to the internals during the change-out process. Clearly, it would be advantageous if the compressor was capable of operating efficiently under winter and summer conditions without the need to change the internals. This would only be possible by using stages with significantly wider operating range. NOMENCLATURE / PERFORMANCE PARAMETERS Before discussing range versus efficiency, some nomenclature must be introduced. To facilitate later discussions, various terms will be used relating to compressor components (i.e., impeller, diffuser, return bend, return channel, and guidevane). These components are labeled in the compressor cross-section shown in Figure 1. In this paper, the term stage refers to a combination of one inlet guidevane, an impeller, a diffuser, and a return channel. The term section refers to a combination of stages; i.e., more than one impeller and its associated stationary hardware. Also, blade or vane angles are measured relative to a radial line, and incidence angles are expressed as flow angle minus blade angle. Next, the parameters commonly used to assess compressor operating range must be understood. The first is a compressor s design or guarantee point or points. Typically, when purchasing a new compression unit, the user will typically select one operating condition at which the performance is to be guaranteed by the manufacturer. The supplier will size the equipment based on this guarantee or design condition. In some cases, in an attempt to insure the necessary operating range, an end-user will specify more than one guarantee condition. The supplier will review the range of conditions to be guaranteed and select one (or some arbitrary point within the required flow range) to be the compressor s design flow condition. This condition is usually

2 where the peak efficiency occurs, though depending on the range requirements, the peak efficiency may occur at a higher or lower flow rate than the selected design condition. A typical compressor map is shown in Figure 2. The flow coefficient is along the x-axis and both polytropic efficiency and head coefficient are along the y-axis. The guarantee flow condition is clearly labeled. Two factors limit the overall flow range of a compressor: surge or stall margin and overload capacity. Surge or stall margin limit the compressor s ability to operate at flow rates lower than design, while the overload capacity limits the ability to operate at higher rates. A tremendous number of factors influence both surge/stall margin and overload capacity including operating speed, gas composition/characteristics, and compressor geometry. It is not the intent of this work to discuss all of these in detail, but rather to introduce the limits to operating range. The term stability is often used to refer to a compressor s surge or stall margin. Stability is typically expressed as a percentage: φ des φsurge / stall Stability = 100 % (1) φdes Where: φ des = flow coefficient at design φ surge/stall = flow coefficient at surge / stall Stability is specified for a given compressor speed line and indicates the amount of flow range from design to surge/stall (see Figure 2). The reader will note the use of the terminology surge/stall margin. The reason is that in many cases, the useable operating range is often not limited by true surge but by some form of rotating stall. The various forms of rotating stall can cause unacceptable levels of subsynchronous radial vibration and therefore limit the ability to use the compressor at low flow rates. Turndown is another parameter used to indicate a compressor s ability to run at lower than design flow. Turndown is determined by tracing a constant head, pressure ratio, or discharge pressure line from design flow back to the surge line (see Figure 3). Like stability, turndown is typically expressed as a percentage. Unlike stability, turndown is not determined at constant speed but, as noted, at constant head, pressure ratio, discharge pressure, or the like. Since the surge line typically has a positive slope, turndown for a compressor will be greater than stability. Rise-to-surge is a way of expressing how much more head or pressure a compressor generates at the surge line as compared with the head or pressure levels at design (see Figure 2). Again expressed as a percentage, rise-to-surge is a key parameter in determining a compressor s or compressor section s controllability; i.e., the greater the rise-to-surge, the easier it is to determine where a compressor is operating on the performance map. Overload capacity and choke margin are terms often used to describe a compressor s ability to operate at higher than design flows. As seen in Figure 2, these parameters reflect the amount

3 that flow may be increased before reaching the maximum allowable flow. Overload capacity is a bit more difficult to define than surge margin since it depends heavily on the supplier s (or user s) interpretation of what constitutes overload. Most compressor manufacturers establish their overload limit based on a variety of considerations such as the following: 1. the drop in efficiency level from design; e.g., -10 points 2. the drop in head level from design; e.g., 30% of design point head level 3. the inlet relative Mach number at the impeller leading edge 4. some minimum allowable efficiency level agreed to by the manufacturer and user Because of the somewhat arbitrary nature of the term overload, it is important that the manufacturer and end-user agree on a common definition. The term range ratio is defined as the ratio of overload flow limit divided by the flow rate at surge for a given speed line (see Figure 2). This parameter has gained wide acceptance amongst purchasers of pipeline boosters. Because range ratio depends on the definition of overload capacity or overload limit, a common understanding between supplier and user must be reached; again, the manufacturer and end user agree to set the overload limit based on some minimum efficiency level. That is, the overload point is established when the efficiency drops to some agreed upon level; e.g., 70% in Figure 2. Having addressed the parameters used to define range, the most common efficiency term used by compressor manufacturers and/or users is polytropic efficiency. The equation is given below: k -1 ln(pr) η p = (2) k ln(tr) where: k = ratio of specific heats Pr = pressure ratio Tr = temperature ratio It should be noted that Equation 2 is only valid for a thermally perfect gas. The expression used to determine the polytropic efficiency for a real gas involves more parameters and is therefore more complex than Equation 2 (ASME PTC10, 1997). The ability of a compressor stage or section to generate pressure is typically expressed as pressure ratio or head rise. Pressure ratio is intuitively obvious and the equations for head and head coefficient, µ P, are given below: Head p η g = p c ( U C - U C ) 2 U2 1 U1 (3) Head gc µ = (4) p P 2 U 2 Where: η p = polytropic efficiency

4 g c = gravitational constant C U1 = tangential velocity of gas entering impeller U 1 = peripheral velocity of impeller leading edge C U2 = tangential velocity of gas exiting impeller U 2 = peripheral velocity of impeller trailing edge If one wishes to calculate the overall head generating capability of a compressor or compressor section, one must determine the head generated by the individual stages in the section or machine and then sum those values to determine the overall head generated. Importantly, all of the parameters described above are used to describe individual stage performance as well as overall compressor or compressor section performance. DESIGN OBJECTIVES Having been active in the pipeline booster market for over 60 years, Dresser-Rand recognized the need to focus on increased range to provide optimal value to clients. Although the OEM s efficiency levels for such units were best-in-class, it was felt that further improvement in overall flow range was possible. Therefore, the focus of the design effort reported in this work was to develop a line of pipeline booster stages that provided improved flow range ratio (range ratio 2.0) while maintaining competitive efficiency levels. DESIGN & ANALYTICAL RESULTS Though a large number of papers have been published in the past 20 years on the subject of 3-D computational fluid dynamics (CFD), centrifugal compressor designers still rely heavily on bulk flow (1-D) and streamline curvature (2-D or quasi-3-d) codes. For example, see references 1 through 4 in the Bibliography. Such tools remain a key part of the design process (see Figure 4). 1-D and 2-D codes execute in a matter of seconds or milliseconds on an average personal computer (PC). Therefore, hundreds of design options can be assessed in a matter of a few hours before embarking on more time-consuming CFD analyses. Before proceeding, it is not the purpose of this paper to describe how these codes function or to discuss the algorithms or solution techniques involved. There is an abundance of literature available that provides such details if the reader desires more information. Bulk Flow (1-D) Analysis Bulk flow design or analysis codes are typically used to establish the basic design parameters for centrifugal stages. Such codes perform calculations at the various critical geometric locations within a stage such as the impeller inlet and exit, diffuser inlet and exit, return channel inlet and exit (if a multistage compressor is being developed) or the volute inlet and exit. Often called velocity triangle codes, these tools use flow through area and/or angular momentum relationships in combination with empirically based loss models to calculate stage performance parameters and bulk flow velocities. Some codes are very simplistic, relying on correlations of specific speed or flow coefficient versus test efficiency to establish the stage efficiency. Others include performance models for individual stage components (impeller, diffuser, return channel, volute, etc.) and then combine the component performance levels to determine the stage levels.

5 Despite their simplicity, 1-D codes aid designers in assessing the performance of a new stage. There are several parameters calculated by 1-D codes that provide guidance to designers with respect to stage operating range. These include local Mach numbers, impeller relative velocity ratio (or diffusion factor), incidence levels in the impeller or stationary components (i.e., the difference between the flow angle approaching the blade or vane and the blade or vane setting angle), and the flow angles themselves. When developing a new impeller, it is important to pay careful attention to the relative velocity ratio, Ws1/W2, where: Ws1 is the relative velocity of the gas along the impeller shroud near the leading edge and W2 is the relative velocity of the gas at the impeller exit. Typically, as Ws1/W2 increases, surge or stall margin decreases. Further, as the impeller design flow coefficient increases, the maximum value of relative velocity ratio must decrease to maintain optimal flow range. In other words, to maintain sufficient flow range, the level of diffusion attainable in high flow coefficient stages must be kept lower than in low flow coefficient stages. Obviously, local Mach numbers can provide insight into the expected range of a stage. As the Mach number increases (as the gas velocity approaches the speed of sound), the ability of a stage or component to pass additional flow decreases. Therefore, if one wishes to maximize the overall range of a stage, it is important to keep the local Mach number as low as possible. 1-D codes can provide designers with good estimates on the incidence angles at the impeller leading edge or at the various other leading edges on the stationary components. For clarity, incidence angles are defined as the difference between the angle of the flow approaching a blade or vane and the blade or vane setting angle (see Figure 18). It is well known that as flow incidence increases, the losses within a component increase. Therefore, one can assess the variation in incidence with flow rate (overload to surge) to determine the sensitivity of a configuration to off-design operation. Quite often, the flow angles themselves provide insight into flow range. For example, it is common knowledge that the onset of diffuser rotating stall (a pressure non-uniformity that forms in the diffuser and limits stable operating range) is a strong function of the flow angle in the diffuser. By comparing the flow angles calculated by the bulk flow models with the numerous widely accepted stall criteria, the analyst can determine the likelihood of stall formation in a proposed design. Regarding the new booster stages, some of the key 1-D parameters are presented in Table 1. The suggested limits for each parameter to insure a reasonable amount of flow range are also shown. As can be seen, the key parameters are within the recommended range to achieve a reasonable trade-off for range versus efficiency. NOTE: The limits suggested in the table may not reflect the experience of other centrifugal compressor manufacturers. However, they do represent generally accepted ranges for high-performing stages [1, 2, 3, and 4]. Streamline Curvature (2-D) Analysis Once an acceptable configuration is developed in the 1-D domain, compressor designers typically will analyze new components or stages using a 2-D or streamline curvature code. Such codes break the aero flow path into streamtubes or stream sheets (see Figure 5) and then use

6 the area distribution and local curvatures within these tubes or sheets to calculate the gas velocity. The most common codes perform detailed calculations in the hub to shroud direction and then apply approximation techniques to estimate the blade-to-blade distributions. The two primary outputs from a streamline curvature analysis are velocity and loading distributions, which are shown for one of the new impeller designs in Figures 6 and 7, respectively. The velocity distribution shows the relative velocity of the gas along the pressure surface and suction surface of the blading within the hub and shroud streamtubes. The loading diagram shows the difference in suction and pressure surface relative velocities normalized by the mid-pitch relative velocity. The loading, often referred to as W/W, is calculated using the following equation: (Wss Wps) / Wm (5) Where: Wss = relative velocity along the suction surface Wps = relative velocity along the pressure surface Wm = relative velocity along the mid-pitch Analysts typically review the velocity and loading diagrams for the streamtubes along the hub and shroud. Since streamline curvature codes have been in use since the early 60 s, a large number of widely accepted assessment criteria have been developed. In addition, most compressor aerodynamic engineers have developed their own specialized criteria based on their positive and negative design experiences. Some of the more commonly held criteria are noted below: Velocity level along the hub streamtube must not drop below zero. Suction surface velocity distribution must not contain regions of rapid deceleration Maximum loading along the hub should not exceed 1.0 Maximum loading along the shroud should not exceed 0.8 In reviewing Figures 6 and 7, one can see that the velocity and load diagrams do adhere to the criteria provided above and generally conform to all accepted guidelines for 2-D impeller analyses. Through appropriate attention to the 2-D analytical results, one can save considerable time on the more complex 3-D CFD analyses described in the following section. 3-D Computational Fluid Dynamics (CFD) Much has been written about the advances of computational fluid dynamics (CFD) and its tremendous value as an analysis tool. The accuracy and/or dependability of the codes have increased markedly during the past few years and many designers / analysts now use the codes as virtual test rigs for new designs. While there are no universally agreed upon acceptance criteria for 3-D CFD results, there are numerous commonsense guidelines that can be followed, especially considering the flowfield details available via CFD. For example, when reviewing velocity vector plots for a stage or

7 component operating at or near design flow, the flowfield must be free of large vortices (other than the expected channel vortex), areas of recirculation, or reverse flow. Such anomalies suggest higher losses and/or reduce operating range. Examples of velocity vector plots are given in Figures 8 and 9. The distributions in Figure 8 are for one of the new booster impellers, while the plots in Figure 9 are for the entire stage. As can be seen, both are free of the adverse situations noted above. For those interested, the compressor was modeled using sector passages (pie slice) of the axisymmetric portion of the main inlet, the three impellers and their downstream diffusers and return channels. The discharge volute and the non-axisymmetric portion of the main inlet were not included. The overall size of the CFD model was about 500,000 nodes. The analysis was performed using CFX-TASCflow assuming steady-state conditions with a frozen rotor interface model between rotating and stationary components. For those unfamiliar with the term frozen rotor, it is an interface scheme applied in CFD codes to make the transition from the rotating frame of reference (the impeller) to the stationary frame (the diffuser or guidevane). The interface assumes that the blades of the impeller and the vanes of the guidevane or diffuser remain in the same relative position throughout the analysis. Though not rigorously correct, if used properly, this approach still provides meaningful insight into the flowfield at the interface between the rotating impeller and the adjacent stationary components. One can also assess the uniformity of the velocity distribution in the hub-shroud or circumferential directions via CFD. Though some variation is natural due to local curvatures, etc., highly skewed flowfields in the hub-shroud direction or non-uniform circumferential distributions may also suggest poor efficiency or reduced flow range. This latter cause came into play during the development of the new booster stages. High flow coefficient impellers typically produce non-uniform velocity profiles in the hub-shroud direction. This is primarily due to the high curvature, short blade length, and reduced radius change along the shroud as compared with the relatively longer blade length, lower curvature, and greater radius change along the hub. A typical impeller exit velocity distribution might look like that shown in Figure 8b, with more tangential velocity along the shroud and more radial velocity along the hub. Such a distribution can influence diffuser performance and could lead to premature stall or reduction in overload capacity. Many novel vaned diffusers have been developed that have proved to be less sensitive to, or have the ability to counteract the hub-to-shroud skewing. These include low solidity vaned diffusers (LSD s), tandem LSD s, twisted or 3-D diffuser vane shapes, and rib diffusers. Numerous technical papers have described the merits of these new diffuser styles including Senoo et al [5] Sorokes and Welch [1992], Amineni et al [1995, 1996], Japikse [1996], and Sorokes and Kopko [2001]. Given the impeller exit velocity distributions, designers selected three styles of diffusers for the new stages vaneless, LSD s, and rib. The initial builds of the new booster stages would include vaneless or low solidity vaned diffusers (LSD s) so that the difference between the vaneless LSD combination and vaneless rib combination could be established. CFD analyses on the new stages suggested that the vaneless diffusers and LSD s would provide excellent pressure recovery at or near design flow conditions. However, there were some

8 indications in the CFD results that the LSD s might limit the overall flow range of the stages (see Figure 10). Still, prior experience with similar impellers suggested that the LSD s were suitable for the new stages. As will be seen later, testing proved that the trends observed in the CFD results were correct. PERFORMANCE TESTING Having addressed the analytical work done to develop the new stages, the discussion will now turn to the performance testing done to validate the new designs. In the past, such validation tests would have been conducted in one of the OEM s single-stage test rigs. These rigs were described in detail in Sorokes and Welch (1992) and Sorokes and Koch (1996). However, one drawback to such rigs is that they do not fully replicate the conditions a stage will experience in a multi-stage configuration; i.e., there is not a full stage upstream. Therefore, the new pipeline booster stages were tested in a multi-stage unit similar to those built for clients. Extensive details on the multi-stage test vehicle are presented in the work of Gilarranz et al (2004), so only a brief overview of the test vehicle will be presented here. Test Vehicle The unit was equipped entirely with the new impellers and stationary components. The first and second stages were designed with vaneless diffuser configurations, while the third stage had an LSD with a two-dimensional profile. Each stage was instrumented to measure the total pressure and total temperature at the impeller eye and return bend. This permitted a performance assessment of each stage individually (see Figure 12). Three sets of probes were distributed circumferentially at each measurement location to ensure redundancy and reduce the uncertainty of the measurements. Combination pressure and temperature ( combo ) probes were used to reduce obstructions in the gas flowpath. Special Instrumentation In addition to the combo probes, four 5-hole pressure probes were located in the second stage of the test vehicle. Two 5-hole probes were placed at the diffuser entrance and two were located at the return bend (see Figure 13). The second stage was selected as best representing the true flow conditions for a multistage compressor since it was an intermediate stage (i.e. there was one stage upstream and one downstream). Selecting the second stage of the compressor also minimized any non-uniform effects caused by the compressor inlet or discharge volute. The 5-hole probes used for this development test were L-shaped and had a hexagonal body (see Figure 14). Knowledge of the probe tip orientation was critical to obtain accurate flow angle measurements. The hex body was used to align the probe during calibration; therefore, the exact position and orientation of the hex of the probe, and hence the tip, was known with sufficient accuracy. Since the test compressor is horizontally split, the probes were placed along the splitline for easy access (see Figures 15 and 16). Slots were machined parallel to the split-line to accurately orient the probe tips at the desired angle. The 5-hole probes at each measuring station were placed at two immersion depths to determine the variations in gas velocity and flow angle at the different depths. The diffuser entrance and return bend each contained one hub side and one shroud side probe (see Figures 15 and 16).

9 The internal instrumentation (5-hole and combination probes) leads were routed out of the compressor through high-pressure seals installed at the end of the machine. The 5-hole probe pressure leads were connected to true differential pressure blocks to obtain high-resolution measurements of the differential pressure between the individual ports on the probe tips. The ability to resolve flow angles from a 5-hole probe is based on pressure differences between the individual ports of the probe. Therefore, it is critically important that the port pressures are measured as accurately as possible. The pressure at the central port (port #1) of each probe was used as a reference value for the measured pressure from the remaining four ports of the 5-hole probe. That is, the pressure measurement for each port located at the periphery of the probe tip was referenced to the pressure at the central port. With this arrangement, the pressure blocks were able to resolve pressure accurately enough to make the 5-hole probe measurements meaningful. The pressures from the central ports of each probe were also acquired by the main data acquisition (DAQ) system so that the pressures of all of the pressure ports of the 5-hole probes were available for data reduction purposes. Description of Test The compressor was operated on a closed loop at the Dresser-Rand test facility in Olean, New York. A Type 2 test [14] was performed to establish the performance of this compressor. Three speed lines were run for the test, each at a different machine Mach number (U 2 /Ao) to encompass the expected operating range of the compressor (U2/Ao = 0.25 to 0.75). Nitrogen was used for the low and design machine Mach number testing, and carbon dioxide was used for the higher machine Mach number testing. In addition to the internal instrumentation, total temperature and total pressure measurements were taken at the inlet and discharge flanges of the compressor. The volumetric flow rate was measured using an orifice plate. The inlet and discharge flange data were used to determine the overall performance of the machine. Initial Test Results As seen in Figure 17, the vaneless LSD build fell short of the goals for both overall flow range and efficiency. The first- and second-stage diffusers were of the vaneless type, while the third stage was equipped with an LSD. The inter-stage instrumentation indicated that the third stage was the root cause of the performance shortfall, limiting both surge margin and overload capacity as suggested by the CFD results referenced earlier. Though the 5-hole probes were installed in the second rather than the third stage, flow angle measurements obtained from the second stage would shed light on the potential problems in the third stage. All impellers in the test vehicle were derived via contour or capacity trimming of a progenitor impeller; i.e., all impellers were from the same impeller family. It is well known that impellers within a given family will have similar characteristics such as head coefficient, efficiency, and exit flowfield. A schematic of the exit velocity triangle with reference to an LSD vane is given in Figure 18. Note that as the volumetric flow rate (Q) through the impeller increases, the magnitude of the relative exit velocity (W2) also increases, causing the flow to become more radial. Clearly, for a fixed geometry LSD vane, optimum incidence cannot be maintained across the full range of operation.

10 The measured flow angles at the second stage diffuser entrance are plotted as a function of flow coefficient (Q/N) in Figures 19 and 20. The incidence angles were derived by subtracting the LSD vane setting angle from the measured flow angle. The hub-side incidence angles are given in Figure 19 while the shroud-side incidences are given in Figure 20. These figures show data acquired at three different values of machine Mach number (U2/Ao). As seen in the figures, the flow angle steadily increases as the flow is decreased from choke to surge/stall. Similar trends are observed in the return bend flow angles as can be seen in Figure 21, which provides a comparison between the flow angles measured at the hub versus the shroud side for the design machine Mach number (U2/Ao). A similar comparison for the diffuser entrance measurements is given in Figure 22. Note that as the machine approaches surge/stall, there is a significant difference between the flow angles along the hub and shroud. This agrees with observations from the CFD results in which the flow exiting the impeller is non-uniform; i.e., more radial close to the hub and more tangential close to the shroud. Further, both the test data and the CFD results suggested that the LSD vanes were limiting the overload capacity. In short, the LSD vanes were not the optimal style for this application. As can be seen in Figure 22, the setting angle of the LSD in stage three is relatively close to the flow angle measured at the shroud side of the machine at the design flow in stage two. Furthermore, at stall conditions, the incidence exceeds a widely accepted limit for incidence on vaned diffusers. THE ANALYTICAL RESULTS REVISITED As noted previously, there were indications in the CFD results of a skewed flowfield exiting the impeller, which could be causing adverse flow incidence on the LSD vanes. A vector plot of the predicted impeller exit velocity for the design operating conditions (flow, speed, etc.) is given in Figure 23. The CFD-predicted impeller exit flow angle distribution is given in Figure 24. The CFD results were taken at the same radius as the 5-hole probes. This figure was also generated for the same operating conditions as figure 23. Both the CFD and test results show a relatively large variation in the impeller exit flow angle from hub to shroud. A comparison of the CFD predicted versus 5-hole probe measured flow angles is given in figure 25. One would not expect perfect matching between the CFD and measured data because of the effects of turbulence and interface modeling as well as the use of a coarse grid density. Still, the trends observed validated the earlier analytical observation that rib diffusers would be a better choice than LSD s for this application. For the subsequent build, the LSD was replaced with a rib diffuser. The rib vanes were mounted on the shroud side. The setting angle of the rib vanes also was adjusted slightly to better match the measured flow angles. Based on prior experience, the expectation was that the rib diffuser would yield increased overload capacity and surge margin. These expectations were supported by CFD results conducted on the rib versus LSD diffusers, which indicated an increase in range with the rib diffusers.

11 FINAL TEST RESULTS A comparison between the performance of the LSD configuration and the rib configuration is given in Figure 26. As can be seen, the rib diffusers yielded a significant improvement in performance. Both the surge margin and overload capacity were dramatically improved and the compressor met the targeted range ratio and efficiency. A comparison of the diffuser entrance flow angles measured at the hub and at the shroud for the original and the modified configuration is given in Figure 27. Note that the flow angles differed little between the two tests, suggesting excellent repeatability of the data measurements. These data also suggest that the diffuser modifications did not cause any detrimental effects on the flowfield exiting the second-stage impeller. In the final analysis, the new impellers met all of the design objectives and the root cause of the stage shortfall was the third stage LSD, as suggested by the CFD results. Once the more appropriate rib diffusers were installed, the performance deficiencies were resolved. In short, the proper mix of analytical work and detailed testing led to an optimized configuration that met the performance objectives. CONCLUSIONS The development effort associated with the new stages yielded some important lessons learned: (1) The intricacies associated with the design of high efficiency, wide flow range stages While it is fairly straightforward to design a stage that delivers either high efficiency or wide flow range, it is more difficult to develop stages that effectively deliver both. The designer / analyst must be keenly aware that customary design guidelines and/or rules and prior experience may be inadequate in these circumstances. (2) Even the most sophisticated analysis tools are not supremely accurate in replicating the complicated flow physics in centrifugal turbomachinery Despite the advances made is recent years, computational models are still numerical approximations of physical conditions. CFD code developers are working closely with engineers from the turbomachinery industry (including the OEM) to improve the accuracy of CFD simulations via advanced turbulence models, discretization schemes and the like. (3) The need to be thorough when reviewing analytical results and to respond properly to trends observed in the results Though computational results can be misleading, they can also provide tremendous insight into the behavior of flow within aerodynamic components. One must review the results with an open mind and not be too quick to discount phenomena that may seem counter-intuitive to the experienced designer. (4) The incredible value in gathering good and detailed data (i.e., flow angles via 5-hole probes) on development stages Good test data will always be more valuable than analytical results in establishing the viability of new stage designs. The value of these data is further enhanced when it can be used to directly validate or calibrate the designer s design / analysis tools such as the case with the flow angles measured by the 5-hole probes. The data from these probes were instrumental in determining the root cause of the performance shortfall.

12 (5) Most importantly, the testing enhanced the application guidelines for LSD s and rib diffusers. LSD s continue to be the diffuser of choice for lower flow coefficient stages with their more uniform exit flow profiles. However, as the flowfield exiting the impeller gets more nonuniform, it is clear that rib diffusers are the more effective alternative. (6) Finally, the testing validated the use of CFD as an effective design filter for determining whether LSD s or rib diffusers are more appropriate for a given application. That is, there were indications in the CFD analyses that rib diffusers would be more appropriate for the new stages and the test data validated those results. It is critically important that all lessons learned are fed back into the OEM s design and analysis systems to improve the quality of design, analyses, and tests conducted during future development efforts. In this case, analytical guidelines for the assessment of 2-D and 3-D results were revised and refined guidelines for the application of rib diffusers versus LSD s were established. Furthermore, techniques were developed to implement 5-hole probes in production equipment, facilitating the use of such probes for future test programs. In the final analysis, the overall design and test program was successful in achieving the OEM s goal of increased range and efficiency. More importantly, the blend of analytical work and test data led to the development of improved design, application and analysis guidelines that will serve the OEM well on future stage development efforts. DISCLAIMER The information contained in this paper consists of factual data and technical interpretations and opinions which, while believed to be accurate, are offered solely for informational purposes. No representation or warranty is made concerning the accuracy of such data, interpretations and opinions. ACKNOWLEDGEMENTS The authors acknowledge the contributions of Mr. Chuck Dunn, Mr. Steve Wilcox, and Mr. E. Bruce Osgood from the Dresser-Rand Test Department for their help with the instrumentation, data acquisition and data reduction. The authors also thank Mr. Jay Koch, Mr. Rob Kunselman, and Mr. Kevin Majot for their assistance in generating some of the figures used herein. Finally, the authors thank Dresser-Rand Company for funding this work and granting permission to publish the results. BIBLIOGRAPHY [1] Aungier, R. H., Centrifugal Compressors, ASME Press, 2000 [2] Cumpsty, N. A., Compressor Aerodynamics, Longman Scientific & Technical, 1989 [3] Japikse, D., Centrifugal Compressor Design and Performance, Concepts ETI, Inc., 1996 [4] Shepherd, D. G., Principles of Turbomachinery, MacMillan Publishing Co., 1956 [5] Senoo, Y., Hayami, H., Ueki, H., Low-Solidity Tandem-Cascade Diffusers for Wide Flow Range Centrifugal Blowers, ASME paper no. 83-GT-3 (1983)

13 [6] Osborne, C. and Sorokes, J.M., The Application of Low Solidity Diffusers in Centrifugal Compressors, Flows In Non-Rotating Turbomachinery Components, ASME FED 69 (1988). [7] Amineni, N., Engeda, A., Hohlweg, W., Boal, C., Flow Phenomena in Low Solidity Vane Diffusers of an Air Packaging Compressor, ASME paper no. 95-WA/PID-1 (1995) [8] Amineni, N., Engeda, A., Hohlweg, Direnzi, G., Performance of Low Solidity and Conventional Diffuser Systems for Centrifugal Compressors, ASME paper no. 96-GT- 155 (1996) [9] Sorokes, J. M., and Koch J. M., 1996, The Use of Single and Multi-Stage Test Vehicles in the Development of the Dresser-Rand DATUM Compressor, Dresser-Rand Technology Journal, 2, pp [10] Sorokes, J. M., and Welch, J. P., 1991, Centrifugal Compressor Performance Enhancement Through the Use of Single-Stage Development Rig, Proceedings of the 20th Turbomachinery Symposium, Texas A&M, pp [11] Sorokes, J.M., and Welch, J. P., 1992, Experimental Results on a Rotatable Low Solidity Vaned Diffuser, ASME Paper 92-GT-19. [12] Benvenuti, E., 1978, Aerodynamic Development of Stages for Industrial Centrifugal Compressors. Part 1: Testing Requirements and Equipment Immediate Experimental Evidence, ASME paper 78-GT-4 [13] Kotliar, M., Engstrom, R., and Giachi, M., 1999, The Use of Computational Fluid Dynamics and Scale Model Component Testing for a Large FCC Prototype Air Compressor, Proceedings of the 28th Turbomachinery Symposium, Texas A&M, pp [14] ASME, 1997, PTC 10, Performance Test Code on Compressors and Exhausters, ASME Press. [15] Gilarranz, José L., Ranz, Andrew, Sorokes, James M. and Kopko, Jason A., On the Use of 5-Hole Probes in the Testing of Industrial Centrifugal Compressors, ASME Paper 2004-GT [16] Sorokes, James M., Kopko, Jason A., Analytical and Test Experiences Using a Rib Diffuser in a High Flow Centrifugal Compressor Stage. ASME paper 2001-GT-320.

14 TABLE 1 Select 1-D Parameters Parameter Recommended Range Actual Value Relative Velocity Ratio, 1.8 (for this impeller flow coefficient range) 1.4 Ws1/W2 Impeller Tip Mach Number, 0.7 (for the target flow range) 0.62 U2/Ao Vaneless diffuser flow angle at 76 (per stall criteria for this configuration) 68 design flow Impeller leading edge incidence Shroud Mean Hub 0 < incidence < < incidence < < incidence < Return channel incidence at design flow 0 < incidence < +5 +4

15 Return Bend Return Channel Inlet Diffuser Volute Inlet Guide Impeller Figure 1 Compressor component nomenclature Range Ratio Head Coefficient Surge Line Rise To Surge Stability Overload Design Point Efficiency Head Coefficient Efficiency Flow Coefficient Figure 2 Performance nomenclature

16 Surge Line Design Point Head Coefficient Constant Head Turndown 100% Speed 90% Speed Head Coefficient Flow Coefficient Figure 3 Illustration of turndown 1-D Design/Analysis 2-D Design/Analysis 3-D Design/Analysis Rig or Production Testing Feedback Loops Figure 4 The typical design cycle

17 Figure 5 Impeller flowpath divided by streamlines Figure 6 Impeller relative velocity distribution Figure 7 Impeller Loading distribution

18 (a) (b) Figure 8 Impeller CFD analyses velocity vectors near mid-span (a) & impeller exit (b) Figure 9 Stage CFD results velocity vectors near mid-span

19 Figure 10 Effects of adverse incidence on LSD vanes during overload operation Figure 11 Instrumentation schematic

20 Figure 12 Combination total pressure / temperature probe Figure 13 5-hole probe locations

21 Figure 14 Schematic of 5-hole probe Figure 15 5-hole probes near hub surface Figure 16 5-hole probes near shroud surface

22 Figure 17 Vaneless - LSD diffuser test results Figure 18 LSD incidence for varying flow rates (a) impeller exit velocity vectors; (b) effect on LSD incidence

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27 Figure 26 Rib diffuser test results (blue) versus vaneless-lsd results (red)

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