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Thinking Outside the Bottle!
The Use of Performance Augmentation Networks to Increase Compressor Efficiency
2011 GMRC Gas Machinery Conference – Nashville, TN – October 3‐5, 2011
W. Norman Shade, PE, ACI Services Inc.
Glen F. Chatfield, OPTIMUM Power Technology
ABSTRACT Intake and exhaust manifold systems are integral parts of modern high‐performance reciprocating engines, enabling major increases in efficiency, specific power output and emissions control. In contrast, the intake (suction) and exhaust (discharge) systems of compressors are generally an after‐thought having the primary purpose of reducing pressure pulsations to tolerable levels. Compressor “manifold” systems reduce pulsations by adding damping that causes pressure losses, reducing the overall compression system efficiency. Improperly designed pulsation control systems increase the adiabatic horsepower required from the compressor cylinders, especially on higher speed compressors. Research over the last five years has explored the development of compressor suction and discharge systems that reduce system pressure losses, while simultaneously augmenting and increasing the compressor efficiency and specific flow output. Previous GMC papers have reported the results of computer simulation studies, lab testing, and field testing of reciprocating compressors with tuned performance augmentation networks (PANTM). Simulations and actual tests confirmed that properly configured PANs were effective in controlling compressor pulsations with very little system pressure loss. Field testing, in cooperation with El Paso Corporation, further confirmed that pulsation bottles, choke‐tubes and orifices could be successfully eliminated from reciprocating compressor systems, replaced by PANs that demonstrated excellent pulsation control, acceptable vibration, and 60 to 80 percent less system pressure loss on low ratio, variable speed pipeline applications. While earlier papers reported PAN developments that utilized a combination of wye‐branches and delay loops to control pulsations with reduced system pressure losses, this paper focuses on more recent research with advanced compressor manifold systems that (similar to modern high‐performance engines) tune the entire compressor system to increase the suction pressure when the cylinder’s suction valves are open and reduce the discharge pressure when the discharge valves are open. This provides a major breakthrough that reduces the required adiabatic horsepower by as much as 15 percent, in addition to eliminating more than 90 percent of the typical pulsation control system pressure losses. For low compression ratio applications, the combined horsepower reductions with PAN systems can reach 30 percent compared with conventional bottle/choke‐tube/orifice systems. In addition to briefly recapping the evolution of this technology, this paper describes in detail the PAN design, simulation, predicted performance and retrofit process for a 6‐throw, 8000 HP, single stage compressor on a major pipeline. Snapshots of computer simulations are used to demonstrate how the technology is applied, and the advanced design processes for optimizing the performance of an entire compression system over a wide range of speed, pressure and unloading conditions.
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INTRODUCTION
Previous papers and articles have discussed the many problems that are commonplace when conventional pulsation attenuation bottle systems are installed on high‐speed (>600 rpm) reciprocating compressors [1, 2, 3]. GMC papers in 2007 [4], 2008 [5] and 2009 [6] tracked the development of a new technology for controlling compressor pulsations and eliminating system pressure losses from initial computer simulations, to successful lab testing, through high‐pressure proof of concept validation testing at a field installation. Recent developments have focused on single‐stage, multi‐throw (4 or 6) high‐speed pipeline compressor systems, where pulsation control, reducing system pressure losses and improving efficiency are particularly challenging problems. These compressor configurations lend themselves to excellent pulsation control, near zero pressure drop losses and significantly improved adiabatic pumping efficiency using only a PAN network of pipes and junctions, with no bottles, choke tubes, orifice plates or even PAN tuned loops. PAN technology makes intelligent and innovative use of pipes, manifolds, junctions and tuned loops, in that order, to improve the performance of reciprocating gas compressors. OPTIMUM’s Virtual Pumping Station simulation and design software was used to predict and compare the pulsation, pressure losses and operating efficiency of specific PAN and bottle pulsation control systems. With ongoing support from the El Paso Corporation, the authors’ companies have collaborated to design a new variable speed and variable load PAN for a large 6‐throw compressor. The most recent reports [7, 8] documented the successful use of a wye‐junction and tuned loops to cancel pulsations without the pressure drop created by traditional pulsation bottles, choke tubes, baffles and orifices. In this paper, the simulated pulsation attenuation and pressure loss performance characteristics of a complete performance augmentation network (PAN) are compared in detail to a conventional complex bottle system for a specific 6‐throw, single‐stage compressor. Further development of the PAN technology has demonstrated the effectiveness of tuning the suction and discharge piping of reciprocating compressors for improved adiabatic pumping efficiency. This paper introduces and explains the theory behind this concept, and the simulated volume flow rate, power consumption and efficiency (BHP/MMSCFD) of the PAN are compared to a typical bottle system. Simulated pressure vs. volume (P‐V) diagrams and crank angle based flange pressure traces provide insight into why bottle systems degrade pumping efficiency. Further simulations show how PAN systems can be tuned to enhance pumping efficiency far beyond the previously reported demonstrated benefit of eliminating essentially all the pressure losses in the system piping. 6‐THROW COMPRESSOR STUDIES Three compressor applications have been studied extensively. The first was an existing 720 rpm constant speed, 8000 bhp electric motor driven, 8.5” stroke Ariel JGV/6 compressor with 18” diameter cylinders. The suction side of the conventional pulsation control system with primary and secondary bottles employing internal choke tubes, baffles and orifices is shown in Figure 1. This bottle system controls pulsations effectively. However, its pressure losses increase the horsepower required to drive the compressor at lower pressure ratios and limit the compressor’s throughput at the higher pressure ratios because unloading is required when the motor’s rated power is reached. A “bottleless” PAN
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system was designed and thoroughly modeled for this installation to control pulsations, virtually eliminate pressure losses and improve efficiency over a wide range of loads at constant speed.
Fig. 1: Existing 6-throw 8000 BHP high-speed compressor showing primary and secondary suction bottles. The second detailed PAN design study was for a similar compressor configuration with no unloading but designed to operate from 750 to 550 rpm (a 27% speed turn‐down). The third and most promising PAN design study, which is discussed in detail below, is for the same size compressor configuration operating under both variable speed and variable load operating conditions. The design specifications called for a 20% speed turn down (720 to 580 rpm) and eight load steps. Many different pressure ratios and load steps were evaluated, with the PAN system showing excellent performance in all cases studied. This third study is the basis for the results that follow. DESCRIPTION OF THE PAN DESIGN
Fig. 2: Isometric and top views of “bottleless” PAN system for 6-throw 8000 BHP high-speed compressor. The design of a PAN system is an iterative optimization process. When a field retrofit is required, it begins with acquiring extensive site measurements and photographs and reviewing available site installation drawings. Then an initial PAN configuration is designed and simulations are run. OPTIMUM’s Automated Design software and Design Optimizing Expert System (DOES) are used to perform multi‐objective optimization on the virtual pumping station to limit pulsations to acceptable levels and
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maximize station efficiency. The configuration is optimized to generally fit within the space constraints of the site. Once promising simulation results are obtained for the target site, a detailed 3‐D CAD model is developed. This provides actual, rather than approximate, piping geometries that are then resimulated. After additional refinements are made with back and forth CAD design and simulation effort, an optimal PAN system such as the one shown in Figure 2 is developed. Final work centers on thorough mechanical analysis and design of supports, mounts and clamps to ensure that the manifold’s vibratory motion and forces are controlled to acceptable levels. It is not a coincidence that the PAN piping that surrounds a modern high efficiency reciprocating compressor resembles the intake and exhaust piping used on modern high performance reciprocating internal combustion engines. Both systems are designed to productively phase and use the substantial energy of the pulses created by their pistons to improve the efficiency of their operation. Dissipating, i.e. wasting, the energy of compressor pulses with bottles, choke tubes and orifice plates is an antiquated idea and a design practice that can be replaced in many compressor applications. Efficient branch junctions, referred to as Tuning Section Transitions (TSTTM), are critical to the performance of the PAN. A compact 2 into 1 “Y‐Branch” TST was designed to efficiently join the two separate OEM cylinder flanges on each suction and discharge into a single standard 12” diameter pipe connection to form the individual runners of the PAN manifold. This Y‐TST would not be required on cylinders having only one suction and one discharge flange. Three 12” header pipes merge the flow and pulses from the three cylinders on each side of the compressor into a “W‐Branch” TST. This was done for both the suction and the discharge for each side of the compressor. Within the W‐TST, each individual pipe area is blended carefully and efficiently into a single 20.35” inside diameter. With the phasing of the crankshaft throws unaltered and fixed, the lengths and diameters of the header pipes are optimized to cancel the pulses joined together at the W‐TST. This cancellation effect is made more effective by unloading all ends of all cylinders equally. At the 20.35” diameter junction of the three header pipes within the W‐TST is a transition nozzle that gradually reduces the inside diameter from 20.35” to 15” so that standard 16” pipe can be attached. As shown in the right‐hand graphic in Figure 2, the discharge PAN system required only a slight modification of the existing pit area on each side of the unit to efficiently turn the flow in the proper direction. Standard pipe and elbows are used for all of the piping between the W‐TSTs and Y‐TSTs. The 16” diameter pipes on the two sides of the compressor are designed to have different lengths before being joined by Y‐TSTs on both suction and discharge. The different pipe lengths from the two W‐TSTs to the Y‐TSTs are optimized to create a phase delay that provides further pulsation cancellation. The large end of the Y‐TST is 24” pipe, which connects to existing station suction and discharge legs coming into the compressor building from the main headers. All of the TSTs are steel castings that can be welded directly into the piping system. This reduces the number of flanges and bolted joints. Flanges can be added at appropriate locations to facilitate installation and/or maintenance access. The suction PAN system is mounted overhead on a pipe deck as shown in Figure 3. This structure is similar to what is employed in chemical plants and refineries. It is rigidly constructed and mounted on reinforced concrete piers. Thorough FEA modeling ensures that there is adequate stiffness to support and restrain the PAN piping overhead during operation. Sufficient space is provided underneath the pipe deck for accessing the frame top cover, and a monorail can be added for frame maintenance if desired.
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Fig. 3: Isometric views of PAN system with suction pipe deck and part of floor decking in place (right). To obtain the best results with PAN technology, both the head and crank ends of all cylinders should be uniformly loaded and unloaded. This makes the pulsations created by the compressor cylinder ends more uniform, and the PAN manifold piping is more effective at attenuating the combined pulsation from all of the ends of the compressor. In this case, each end of each cylinder was configured with 3 different sized (175, 350 and 700 in3) pneumatically actuated fixed volume clearance pockets. This configuration provides 8 different unloading steps ranging from 0 to 57% added clearance on each cylinder end. This configuration actually provides more effective range and better efficiency than running all 6 cylinders single acting. In addition, the ability to run variable speed between 580 and 720 rpm provides a very large continuous operating range. COMPARISON OF PULSATION CONTROL
Fig. 4: PAN and bottle system suction (left) and discharge (right) pulsation control for 1.2 ratio, full load. Detailed simulation results compare the compressor system performance of the PAN and the bottle systems for the extremes of this compressor’s intended operating range: pressure ratio 1.2 with no unloading and pressure ratio 1.5 fully unloaded. The terms “no unloading” and “fully loaded” mean that the compressor has no volume pockets open and no end deactivations, so that all compressor cylinder
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ends are fully active to generate maximum flow. The term “fully unloaded” means that all volume pockets are open on all cylinder ends to minimize flow at a given speed. Figure 4 compares the simulated suction and discharge pulsation at the pipeline connections of the PAN and bottle systems over the 580 to 720 rpm speed range at a 1.2 pressure ratio with no unloading. In this case the PAN’s maximum suction pulsation was 4.3 psi (0.7% of line) and maximum discharge pulsation was less than 3.4 psi (0.5% of line). Both are well within industry standards. Figure 5 shows the actual simulated pulsations and attenuation for this same case at 720 rpm at the cylinder flanges, after the W‐TSTs and after the Y‐TSTs in the pipeline headers.
Fig. 5: PAN system suction (left) and discharge (right) pulsation control for 1.2 ratio, full load. Figure 6 compares the simulated suction and discharge pulsation at the pipeline connections of the PAN and bottle systems over the 580 to 720 rpm speed range at a 1.5 pressure ratio fully unloaded (all pockets open). In this case the PAN’s maximum suction pulsation was 3.7 psi (0.7% of line) and the maximum discharge pulsation was 4.7 psi (0.6% of line). Again, both are well within industry standards. Figure 7 shows the actual pulsations for this case at 720 rpm at the cylinder flanges, after the W‐TSTs and after the Y‐TSTs on the pipeline headers.
Fig. 6: PAN and bottle system suction (left) and discharge (right) pulsation control for 1.5 ratio, unloaded.
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Fig. 7: PAN system suction (left) and discharge (right) pulsation control for 1.5 ratio, unloaded. COMPARISON OF SYSTEM PRESSURE LOSSES While this PAN is very effective at controlling pulsation, its pressure loss reduction is even more impressive. Figure 8 compares pressure losses from simulations of the existing eight‐bottle system and the PAN system for both the 1.2 ratio loaded and the 1.5 ratio unloaded cases. At the 1.2 ratio, with no unloading, the total of suction and discharge pressure losses of the PAN system at all speeds from 580 to 720 rpm is approximately 1.0 psi. By comparison, the baseline bottle system experiences a total pressure loss of over 9.6 psi at 580 rpm, increasing steadily with increasing speed to 14.5 psi at 720 rpm. At the highest flow at 720 rpm, the PAN has 93% less pressure loss than the bottle system. On the suction side of the PAN, the Bernoulli Effect creates a slight pressure gain, not a loss, because this compressor has a total flange area that is slightly larger than the area of the 24” suction pipeline connection. At the lower flow 1.5 pressure ratio conditions, the PAN pressure loss is 0.3 psi or less. Under the same operating conditions, the current bottle system has a total pressure loss ranging from 4.6 psi at 580 rpm to 8.2 psi at 720 rpm. At 720 rpm the PAN has 96% less pressure loss than the bottle system.
Fig. 8: PAN and bottle system pressure loss for 1.2 ratio, fully loaded (left) & 1.5 ratio, fully unloaded (right). The data shows that the reduced pressure losses in the PAN system result in the elimination of more than 90% of the horsepower consumption required to overcome the pressure loss caused by traditional
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bottle‐choke tube‐orifice pulsation control systems. This translates into a significant system efficiency gain, especially in low pressure ratio pipeline applications. COMPARISON OF COMPRESSOR ADIABATIC PUMPING EFFICIENCY Another significant factor to consider is how the flow rates and adiabatic efficiencies compare between the PAN and bottle systems. Figures 9 and 10 provide a comparison at a 1.2 pressure ratio (suction pipeline pressure 625 psig and discharge pressure 750 psig) with no unloading. Figure 9 shows that at 1.2 ratio the fully loaded volume flow rates of the PAN and the bottle system are essentially the same over the entire speed range (left figure). It also shows that, with the PAN system, the compressor requires far less power to pump the the same amount of gas (right figure). At 720 rpm the bottle system requires more than 1000 additional BHP to operate the compressor at the same flow rate. It should be noted that, although the goal of this PAN design was to keep the flow rate the same, alternative PAN systems can be designed to increase flow through the compressor.
Fig. 9: PAN vs bottle system for 1.2 ratio fully loaded volume flow rate (left) & compressor power (right)
Fig. 10: PAN vs bottle system for 1.2 ratio fully loaded BHP/MMSCFD (left) & PAN % efficiency gain (right) Figure 10 shows the efficiency of both systems measured in BHP/MMSCFD and the % PAN efficiency improvement for 1.2 ratio fully loaded. The PAN’s efficiency is 15% better at 720 rpm and it averages more than 12% over the full speed range. This gain is includes the savings that result from reduced system piping losses.
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Figures 11 and 12 provide a comparison at a 1.5 pressure ratio fully unloaded. Figure 11 shows an interesting synergy between this particular PAN design and pocket unloading. The same pockets provide 15% to 20% more flow reduction (left) or unloading with a PAN than they do with a bottle system. This sharply reduces the need for inefficient cylinder deactivation, which is a souce of excessive pulsation and high vibration. Figure 11 also shows that under these operating conditions the PAN requires significantly (25 to 30%) less power (right), partly due to the more effective unloading and partly due to the reduced adiabatic power required.
Fig. 11: PAN vs bottle system for 1.5 ratio fully unloaded volume flow rate (left) & BHP (right) Figure 12 shows the efficiency of both systems measured in BHP/MMSCFD and the % PAN efficiency improvement for 1.5 ratio fully unloaded. It shows that even under the lowest volume flow conditions, this PAN is about 7% more efficient than the bottle sytem across the entire speed range. This gain includes the savings that result from reduced system piping losses.
Fig. 12: PAN vs bottle system for 1.5 ratio fully unloaded BHP/MMSCFD (left) & PAN % eff. gain (right) The virtual pumping station design and analysis software produces pressure vs. volume (P‐V) diagrams for any head or crank end of a virtual compressor. Figure 13 compares the P‐V diagrams of the same cylinder head‐end for both the PAN and bottle systems. For this case, the suction pipeline pressure is 625 psig and the discharge pipeline pressure is 750 psig resulting in a 1.2 pressure ratio. The area within the curves is proportional to the work that the cylinder end does to pump the gas. Comparison of the envelopes in Figure 13 clearly shows that when connected to the PAN system the cylinder end does less work to pump the same amount of gas and is therefore much more efficient.
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Fig. 13: PAN vs bottle system Pressure vs Volume diagram for 1.2 ratio fully loaded @ 720 rpm The gain in pumping effiency and reduction in adiabatic power are intentional results from the PAN design process that seeks to improve the cyclical pressure fluctuations at the cylinder flanges. In addition to analyzing P‐V diagrams, the designer can obtain gas presure vs. crank angle plots at any pipe location along the gas flow path within the virtual pumping station model. A close inspection of the cyclical pressure fluctuation at the cylinder suction and discharge flanges for both the PAN and bottle systems show why the PAN system requires less power. The actual dynamic animations available in the modeling software make this effect much easier to comprehend than the several screen capture plots that can be included herein, however they serve to explain the concept fairly well. Figures 14 and 15 compare the instantaneous gas pressure vs. crank angle at the suction and discharge flanges, respectively, of the PAN and bottle system at a pressure ratio of 1.2 and 720 rpm. In Figure 14, it can be seen that when the compressor cylinder is drawing gas into both ends, the bottle suction flange pressure with the bottle system is much lower than the suction flange pressure with the PAN system most of the time. While the bottle system suction pressure loss from the pipeline to the flange (Suc Pipeline ‐ Bottle Suc Avg average for 360 degrees) is 9 psi, the dynamic pressure difference between the bottle system and the PAN system during the suction flow event can exceed 30 psi. In Figure 15, it can be seen that the discharge flange conditions are the opposite of the suction. Here the cylinder discharges gas into a higher pressure with the bottle system than with the PAN system. The difference between the bottle system discharge pressure loss and PAN system discharge pressure loss is only 3 psi. However, the difference in flange pressure at mid‐stroke (‐90° and +90°) is 47 psi. The timing of the pressure pulses is the key to reducing the adiabatic power, and this is a vivid example of how averages can be very misleading for dynamic gas systems.
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Fig. 14: PAN vs. bottle system suction flange press. vs. crank angle for 1.2 ratio no unloading @ 720 rpm
Fig. 15: PAN vs. bottle system disch. flange press. vs. crank angle for 1.2 ratio no unloading @ 720 rpm
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CONCLUSIONS AND SUMMARY Conventional bottle systems are designed to dissipate the energy in the pressure pulses that come out of reciprocating compressor cylinders. The pulse energy is treated as a problem that must be eliminated as quickly as possible so that shaking forces are minimized. In managing the pressure pulsation control, bottle diameters tend to get very large, which of course, can increase the dynamic shaking forces that result from the pressure pulsations. A well designed PAN system, however, takes a much different approach. Rather than dissipate pressure pulse energy, a PAN uses that energy to make the compressor operate more efficiently, while simultaneously attenuating, or cancelling, pulsation at the pipeline to very low levels. The design process involves changing the phasing and amplitude of the gas pulsation at the flanges to maximize the cylinder’s adiabatic pumping efficiency. Although this concept has been employed successfully in engines for several decades, it is a new concept for high‐pressure reciprocating compressors that can result in major compression efficiency gains. The research and development accomplished over the past several years shows that the proper application of PAN technology accomplishes four important compressor pumping station objectives.
First, it attenuates, or cancels, the detrimental gas pulsations created by reciprocating compressors, preventing them from traveling upstream and downstream from the pumping station.
Second, a PAN system reduces the magnitude of the vibratory forces compared to pulsation bottles. In PAN systems, pulsations are reduced significantly at the first manifold junction. This means that the high amplitude pressure pulsation acts only on the relatively short, smaller diameter header pipes (12” in this example), rather than the much larger diameter (up to 60”) bottles. It is estimated that the resulting pulsation induced shaking forces can be 25% to 90% lower, which should reduce vibration related failures and increase the mean time between failures for the compressor.
Third, PANs minimize pressure losses on both suction and discharge. There are no baffles, choke tubes or orifice plates to impede gas flow. Instead, the system employs properly sized smooth bore pipes and transition adapters (TSTs) that minimize flow velocity and turbulence.
And fourth, PAN technology can be used to significantly reduce the compression horsepower requirement. Optimized PAN configurations can boost the suction pressure during the compressor cylinders’ suction events and reduce the discharge pressure during the cylinders’ discharge events. Combining this with the aforementioned reduction of system pressure losses can reduce pipeline compression energy costs by 10% to as much as 30%.
ACKNOWLEDGEMENTS The authors wish to acknowledge the El Paso Corporation for providing significant support that has enabled this research to progress over the past five years. Special thanks are acknowledged to Randy Raymer, Tom Burgett and others at El Paso for their encouragement and support for the development and practical adaptation of this technology to the gas compression industry. Finally, the management and owners of Optimum Power Technology and ACI Services Inc. are recognized for their ongoing funding, support and dedication of key technical personnel and other resources to this ongoing research.
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REFERENCES
1. Harris, R. and Raymer, R., The Value ($$) of Compressor Efficiency, 2006 GMC Short Course, Oklahoma City, OK, October 5, 2006.
2. Greenfield, S. D., Optimizing Compressor Design for Complex Reciprocating Installations, 2006 GMRC Gas Machinery Conference, October 2‐4, 2006.
3. Chatfield, G. and Shade, W.; New Technology for the Efficient Cancellation of Compressor Pulsations, GMC Journal, February 2009.
4. Brahler, C.; Chatfield, G.; Crandall, J; and Shade, W.; An Investigation of the Application of Finite Amplitude Wave Simulation with a New Technology for Controlling reciprocating Compressor Pulsations, 2007 GMRC Gas Machinery Conference.
5. Chatfield, G.; Crandall, J.; Shade, W.; and Wells, D.; Demonstration of Efficient Compressor Control Using Tuned Loop Networks, 2008 GMRC Gas Machinery Conference.
6. Bazaar, J.; Chatfield, G.; Crandall, J.; Shade, W. and Wells, D.; Efficient Bottle‐Less Compressor Pulsation Control – Experimental Test Results, 2009 GMRC Gas Machinery Conference.
7. Shade, W.; Efficient Bottle‐Less Compressor Pulsation Control, GMC Today, October 2009. 8. Chatfield, G. and Shade, W.; Thinking Outside the Bottle: Attenuate Pulsation and Eliminate 90% of All Pressure
Losses, GM Journal, May 2011.