Economic Benefits Of Compressor Analysis > Improving gas flow through a compressor maximizes value and revenue By EDWARD B. FLANAGAN, PE Edward B. Flanagan, PE, is the general manager at Windrock Inc. His area of expertise is systems analysis and design, analog and digital design, instrumentation and controls, and the application of instrumentation for machinery health and control. Introduction Large-frame reciprocating compressors have been used widely in oil and gas applications since the 1920s. As the large number of facilities utilizing this type of equipment has increased, so has the need for maximizing operational reliability while minimizing maintenance expenditures. Compressor analysis is a critical tool that provides a health condition assessment, ensures efficient and safe operation, accurately measures machinery performance, and provides a measurable economic benefit for the operator. This paper investigates excessive fuel consumption of compressor drivers caused by common compressor faults. Pressure versus volume (PV) analysis techniques will identify deficiencies, quantify fault severity, and will be used to estimate the resulting excessive fuel consumption. Empirical fuel measurements of the drivers are analyzed before and after the fault correction and are used to calculate immediate economic savings from repairs. Performance and capacity improvements are also analyzed, providing a complete economic picture of maintenance and operational payback. Background Reciprocating compressor analysis has been in use for many years, growing from the oscilloscope-based analyzers to modern digital portable and online systems. These tools have allowed analysts to evaluate compressor health, accurately measure performance and provide protection against potential catastrophic failures. However, the economic benefit of compressor analysis and conditionbased maintenance is often difficult to quantify or is not utilized. The measure of in-cylinder dynamic pressure versus crank angle provides the cornerstone of compressor analysis. Often denoted as pressure versus time (PT), these curves can be converted into a pressure versus swept volume (PV) plot. By comparing actual versus theoretical PV plots, leakage within the compressor cylinder can be identified, including suction/discharge valves, rings, and packing. These leakages are often confirmed and further distinguished by vibration, ultrasonic and temperature readings, as well as numeric data derived from the pressure readings, such as flow balance and leak index. Pressure data are used to calculate compressor performance parameters, such as horsepower consumed, gas throughput, flow and efficiency. Finally, pressure and vibration data play a critical role in identifying potential failure modes, including rod overload, lack of rod reversal, and crosshead looseness. A leak within a compressor cylinder often results in the recirculation of gas. For example, if a discharge valve is leaking during the suction event, gas, which has already been compressed to discharge pressure, will leak back into the cylinder while it is at a lower pressure. Recompressing the same gas results in less throughput and additional power requirements. The amount of throughput or flow can be accurately calculated by thermodynamic analysis of the cylinder PV parameters. The additional cost to recompress the gas can be estimated by examining the ratio of gas moving into the cylinder versus the amount of gas leaving the cylinder (flow balance), and multiplying the percentage of this loss by the horsepower cost for the cylinder to compress gas. The resultant is termed recirculation loss, an estimate of the economic value of the cost of the leak, often expressed in $/day or $/year of compressor operation. While the recirculation loss can provide an estimate of the cost of cylinder leakage, to measure the actual cost, power consumption measurements are necessary. Since many highspeed compressors are driven by natural gas engines, measuring fuel flow and multiplying by fuel cost can
n Figure 1. Case 1: discharge valve leakage. The unit is a Caterpillar 3608 engine driving an Ariel JGD-4 three-stage compressor with a Windrock 6320/AP compressor analyzer. The application for Case 1 is natural gas gathering. n Figure 2. Cylinder 4 PV with leaking discharge valve(s)(top). n Figure 3. Cylinder 4 PT and ultrasonic traces confirming leaking discharge valves (bottom).
calculate power consumption. The result is expressed in $/day. When that number is divided by the actual compressed gas flow, the cost per flow or $/MMscfd is determined. By calculating this value for a compressor with a known valve leak and then again after the valve is repaired, the actual fuel savings are the difference. In addition to excessive fuel requirements, compressor faults often result in loss of throughput. If a compressor is incapable of providing its rated throughput capacity and that capacity is required (to be available for sale or because it limits a larger production process), this loss of flow can directly result in significant economic losses. For example, a natural gas producer may be able to sell gas for US$3.50/MMBtu. Assuming a heating value of 1024 Btu/ scf (38,153 kj/m 3 ), that provides potential revenue of US$3580/MMscfd or US$1,308,000/MMscf/yr. If by repairing a known compressor fault, the producer is able to flow an extra 1.0 MMscfd (0.028 x 10 6 sm 3 /d), the resulting sale of the gas is worth US$1.3 million per year. n Figure 4. Rod load for cylinder 3. Note the gas load is near the maximum tension and compression rating, and there is significant vibration knock at the crosshead. n Figure 5. One of the two faulty discharge valves that were replaced. Compressor Cylinder ID ihp @ rpm ihp (MMscfd) Capacity (MMscfd) Date Time 1 > Comp 1 H Pressure 198.0 @ 991.5 79.71 2.48388 9/8/14 16:01:48 2 > Comp 1 C Pressure 201.1 @ 991.2 79.01 2.54490 9/8/14 16:03:32 3 > Comp 2 H Pressure 314.3 @ 991.0 48.98 6.41673 9/8/14 16:16:55 4 > Comp 2 C Pressure 282.7 @ 992.0 47.40 5.96321 9/8/14 16:18:47 5 > Comp 3 H Pressure 198.7 @ 991.6 79.17 2.50968 9/8/14 16:06:57 6 > Comp 3 C Pressure 199.0 @ 986.9 78.88 2.52300 9/8/14 16:09:10 7 > Comp 4 H Pressure 365.0 @ 990.8 51.13 7.13873 9/8/14 16:12:39 8 > Comp 4 C Pressure 314.5 @ 987.6 45.16 6.96402 9/8/14 16:14:07 % Volumetric Efficiency % Power/Valve Loss Discharge Suction Discharge Suction % Flow Balance Suction/ Discharge Toe Pressure Rod Temperature ( F) Rod Load (%) Minimum Ratio Pd Ps Discharge Suction Tension Reversal 1 > 15.2 50.4 3.6 3.3 1.07 206.12 39.68 4.06 247.8 79.3 97.9 96.9 135 C 2 > 15.9 52.0 3.3 3.7 1.05 205.64 39.26 4.08 247.8 79.3 97.9 96.9 135 C 3 > 40.2 78.5 5.4 4.9 0.99 949.92 398.16 2.34 227.0 113.3 62.1 70.3 127 C 4 > 40.6 78.5 6.4 5.9 1.03 935.21 409.28 2.24 227.0 113.3 62.1 70.3 127 C 5 > 15.3 50.2 3.2 3.0 1.06 207.94 40.25 4.05 251.4 79.3 98.0 96.4 134 C 6 > 15.8 51.5 3.0 3.7 1.07 207.00 40.80 3.99 251.4 79.3 98.0 96.4 134 T 7 > 56.1 72.9 7.1 3.9 0.78 424.03 203.89 2.01 246.0 123.9 62.4 69.4 118 T 8 > 45.6 81.4 7.3 5.3 1.07 411.38 197.24 2.01 246.0 123.9 62.4 69.4 118 T n Table 1. Compressor report before repair. Note the low flow balance for cylinder 4 head end (red box) and the high rod loads for both stage-one cylinders (blue boxes).
n Figure 6. PV after repair. The following two examples illustrate use of compressor analysis to identify faulty conditions and determine the economic benefits of the repairs (Figure 1). Case 1 is based on analysis of the Figure 1 compressor and Case 2b is based on analysis of the Figure 9 compressor. Analysis of the compressor showed significant discharge valve leak on cylinder 4 (stage 2), illustrated by the PV curves in Figure 2. The head-end pressure trace (solid blue curve) does not follow the theoretin Figure 7. Rod load for cylinder 3 after repair. Note that rod load is lower and the crosshead knock is gone as compared to Figure 4. Compressor Cylinder ID ihp @ rpm ihp (MMscfd) Capacity (MMscfd) Date Time 1 > Comp 1 H Pressure 177.9 @ 988.2 74.63 2.38436 9/8/14 18:53:58 2 > Comp 1 C Pressure 186.0 @ 993.2 72.88 2.55269 9/8/14 18:52:55 3 > Comp 2 H Pressure 318.2 @ 991.3 46.36 6.86408 9/8/14 18:44:02 4 > Comp 2 C Pressure 286.4 @ 990.6 44.57 6.42670 9/8/14 18:45:13 5 > Comp 3 H Pressure 184.9 @ 996.0 75.84 2.43855 9/8/14 18:55:13 6 > Comp 3 C Pressure 185.8 @ 990.8 71.93 2.58329 9/8/14 18:56:11 7 > Comp 4 H Pressure 314.0 @ 984.5 53.18 5.90486 9/8/14 18:48:39 8 > Comp 4 C Pressure 308.6 @ 988.6 53.57 5.76106 9/8/14 18:50:16 % Volumetric Efficiency % Power/Valve Loss Discharge Suction Discharge Suction % Flow Balance Suction/ Discharge Toe Pressure Rod Temperature ( F) Rod Load (%) Minimum Ratio Pd Ps Discharge Suction Tension Reversal 1 > 16.7 48.4 2.9 2.6 0.98 181.65 36.96 3.80 236.5 83.4 89.0 86.8 154 C 2 > 17.6 52.4 2.6 3.4 1.00 181.29 36.94 3.80 236.5 83.4 89.0 86.8 154 C 3 > 41.6 79.9 5.3 5.4 1.01 959.75 418.54 2.25 218.6 111.2 69.0 76.4 162 C 4 > 42.9 79.9 6.8 6.5 1.03 941.17 431.14 2.14 218.6 111.2 69.0 76.4 162 C 5 > 16.8 48.4 2.5 2.8 0.97 183.49 37.38 3.81 233.2 80.0 88.6 87.4 154 C 6 > 17.8 52.1 2.7 3.2 1.00 181.22 37.60 3.75 233.2 80.0 88.6 87.4 154 C 7 > 34.8 71.1 6.0 4.1 1.02 440.03 174.37 2.41 238.9 118.9 82.3 85.0 154 C 8 > 34.9 73.6 4.9 3.9 1.04 439.58 171.39 2.44 238.9 118.9 82.3 85.0 154 C n Table 2. Compressor report after repairs. Note the improved flow balance (red box) and reduced rod loads (blue boxes) compared to Table 1.
Parameter Source Before Repair After Repair Difference Compressor Power (hp) Gas Flow (MMscfd) Ratio Fuel Flow (scfh) Windrock MD Software Windrock MD Software Windrock MD Software n Figure 8. Theoretical losses. 2182 2065-5.40% 12.38 13.29 7.35% 17.9 18.7 4.50% Measured 17,285 16,983-1.70% hp/mmscfd Calculated 176.3 155.4-11.90% Fuel Flow/hp Calculated 7.92 8.22 3.80% US$/hp-hr (@ $4/Mscf) Calculated $0.031 $0.032 US$/d Calculated $1,660 $1,630-1.80% US$/MMscfd Calculated $134.10 $122.65 8.5% ($11.45) cal curve (dashed blue line). With the measured pressure coming up to discharge pressure too quickly and back down to suction pressure too slowly, a classic discharge leak pattern is identified. Figure 3 confirms the leak with ultrasonic measurements. As indicated, the ultrasonic measurements should be quiet (thin line) during the head-end suction event. The excess noise in this area is indicative of a discharge valve leak. Also note in Figure 2 that the automated Leak Index tool also identifies the discharge leak. The calculated performance parameters also confirm a discharge valve leak with a low flow balance (below 0.95) as shown in Table 1. The low flow balance indicates that more gas is flowing into the cylinder than is leaving the cylinder over the course of each stroke. The leaking discharge valve in the cylinder 4 head end is allowing discharge gas to recirculate back into the cylinder during the suction stroke. The poor efficiency in stage 2 caused stage 1 cylinders to perform a higher share of the work for compressing the gas. Because of this imbalance, cylinders 1 and 3 are dangerously close to rod overload and have caused a significant vibration knock in the crosshead when the piston changes from tension to compression. The overload can be seen numerically in Table 1 and graphically in Figure 4, which also shows the crosshead vibration knock. With the compressor analysis pinpointing discharge valve leakage on the head end of cylinder 4 and a potential for catastrophic failure of cylinder 3, the maintenance crew opened cylinder 4 to confirm and replace the faulty discharge valves (Figure 5). After the repairs were completed, the actual PV and theoretical PV curves lined up very closely, as shown in Figure 6. The flow balance improved from 0.78 to 1.02 for the cylinder 4 head end, as shown in Table 2. The same table also shows a reduced rod load for the first-stage cylinders. Figure 7 shows the ren Table 3. Compressor performance and fuel flow economics. n Figure 9. Case 2: suction valve blockage. The unit is a Caterpillar 3516 engine driving an Ariel JGT-4 three-stage compressor with a Windrock 6320/PA compressor analyzer. The application for Case 2 is natural gas gathering.
Compressor Cylinder ID ihp @ rpm ihp (MMscfd) Capacity (MMscfd) Date Time 1 > Comp 1 H Pressure 47.8 @ 1382.8 64.77 0.73760 9/14/14 12:46:15 2 > Comp 1 C Pressure 107.0 @ 1381.2 72.46 1.47653 9/14/14 12:47:22 3 > Comp 2 H Pressure 187.2 @ 1382.8 70.18 2.66791 9/14/14 12:52:53 4 > Comp 2 C Pressure 157.1 @ 1382.6 63.34 2.48013 9/14/14 12:54:14 5 > Comp 3 H Pressure 49.5 @ 1380.8 72.08 0.68704 9/14/14 12:50:25 6 > Comp 3 C Pressure 112.2 @ 1381.7 69.93 1.60473 9/14/14 12:51:19 7 > Comp 4 H Pressure 179.1 @ 1378.6 77.46 2.31222 9/14/14 12:55:19 8 > Comp 4 C Pressure 159.4 @ 1380.2 74.86 2.12891 9/14/14 12:55:54 % Volumetric Efficiency % Power/Valve Loss Discharge Suction Discharge Suction % Flow Balance Suction/ Discharge Toe Pressure Rod Temperature ( F) Rod Load (%) Minimum Ratio Pd Ps Discharge Suction Tension Reversal 1 > 11.3 27.0 8.7 4.0 1.04 117.40 29.70 2.98 220.0 68.0 50.7 49.2 126 T 2 > 23.1 57.3 10.0 6.8 1.04 116.50 28.35 3.05 220.0 68.0 50.7 49.2 126 T 3 > 34.1 75.8 9.1 15.7 1.02 877.07 304.80 2.79 241.8 85.0 53.5 57.6 145 T 4 > 35.7 74.2 9.3 13.8 1.04 863.56 326.40 2.57 241.8 85.0 53.5 57.6 145 T 5 > 10.0 24.9 5.7 3.2 1.02 123.70 29.70 3.12 218.2 71.2 50.5 49.7 125 T 6 > 24.9 58.7 9.4 6.2 1.00 117.60 29.41 3.00 218.2 71.2 50.5 49.7 125 T 7 > 29.7 75.7 9.9 7.8 1.04 386.47 108.43 3.26 256.1 79.1 61.0 62.8 137 C 8 > 28.8 74.5 8.7 7.4 1.08 381.39 110.83 3.16 256.1 79.1 61.0 62.8 137 C Stage Capacity (MMscfd) 1 4.5059 2 4.4411 3 5.148 n Table 4. Compressor report before repair. Note the high power losses for the suction events for cylinder 2 (red boxes). duced rod load graphically, as well as the elimination of the crosshead knock that was seen in Figure 4. n Figure 10. PV curves before repair. Toe Pressure Pd Ps Ratio 117.40 29.70 2.98 116.50 28.35 3.05 877.07 304.80 2.79 863.56 326.40 2.57 123.70 29.70 3.12 117.60 29.41 3.00 386.47 108.43 3.26 381.39 110.83 3.16 n Table 5. Cylinder pressures. Economic analysis for discharge valve leak The leaking discharge valves in the head end of cylinder 4 (stage 2) caused discharge gas to re-enter the cylinder during the suction stroke. This gas expands during suction and is recompressed during the compression stroke. Recirculating this gas causes additional work by the driver. Windrock MD software uses the flow balance to estimate the percentage of recirculated gas and multiplies that by the horsepower required for the head end of cylinder 2. Using a standard horsepower cost of $0.032/hphr (based on a fuel cost of $4/Mscf), the software calculated a recirculation loss of $112/d or nearly $41,000/ yr. Figure 8 represents the excess power required to recompress gas that leaked back in through the faulty discharge valves.
By evaluating the fuel flow readings before and after the repair, a more precise economic analysis can be developed, as shown in Table 3. After the repair, the compressor moves more gas with less horsepower. Assuming this compressor is running 24/7 at a flow rate of 13.29 MMscfd (0.376 x 10 6 sm 3 /d), a fuel savings of US$55,500 per year is realized with the replacement of the two head-end discharge valves on cylinder 4 (stage 2). In addition to these fuel savings, there are intangible savings from lowering the rod loading and distributing the load among the stages more evenly, resulting in less component n Figure 11. Clogged suction valves. n Figure 12. PT curves before and after repairs. wear and avoidance of potential catastrophic failure. Initial analysis revealed a relatively healthy compressor with little recirculation. In contrast to the leaking discharge in Case 1, the actual PV plots follow closely with the theoretical curves, the flow balances are close to 1.0 (indicating a healthy cylinder), and the ultrasonic data does not show any cylinder leakages. Upon closer inspection of the compressor report data in Table 4, a higher than normal power loss is found during the suction event for cylinder 2 (stage 3). Power loss represents the energy required to pull the gas into the cylinder from the suction bottle. It can have several components: the power required to open and flow gas through the suction valves and pressure losses through the nozzles and any restrictions, such as orifice plates. As seen in Figure 10, the PV curve shows very high suction horsepower losses, confirming the data in the compressor report. The values of 15.7 and 13.8% suction valve horsepower losses computes to 50 hp (37.3 kw) additional load required to pull in gas to this compressor cylinder. A more typical figure would be 7 or 8%. The cause of this excess power requirement could be clogged suction valves. Another possibility would be a clogged screen in the nozzle of this machine, or even a clogged cooler between the second and third stages. The odd pointed down shape of the PV toe indicates excessive drawing down of the bottle pressure, representative of a blockage between stages. From close examination of the compressor report, there is a pressure drop from the first to the second stage of 7 psi (0.48 bar), and a drop from the second to the third stage of 70 psi (4.83 bar). This is also indicative of a blockage. The compression ratios across the stages are not balanced, resulting in an uneven distribution of the work across the compressor. The third-stage suction valves were pulled and found to be significantly clogged with a white crystalline substance, as can be seen in Figure 11. After the repair, additional analysis data were gathered on the machine to quantify the improvement made by changing the clogged suction valves. Compressor Cylinder ID ihp @ rpm ihp (MMscfd) Capacity (MMscfd) Date Time BEFORE 3 > Comp 2 H Pressure 187.8 @ 1382.8 70.93 2.64709 9/14/14 12:52:53 4 > Comp 2 C Pressure 157.5 @ 1382.6 63.72 2.47112 9/14/14 12:54:14 AFTER 3 > Comp 2 H Pressure 199.9 @ 1382.5 52.55 3.80416 10/28/14 12:28:43 4 > Comp 2 C Pressure 165.5 @ 1379.8 49.88 3.31746 10/28/14 12:29:28 n Table 6. Measured performance data before and after valve repair.
Figure 12 illustrates the pressure waveforms before and after the suction valves were replaced. It is visually apparent from the waveforms that the clog caused a severe restriction and pressure drop inside the cylinder during suction, thus lowering the final discharge capacity of the cylinder. Economic analysis of restricted suction valve From theoretical models, repair of this issue resulted in about a 25 hp (18.64 kw) reduction in load for the same amount of production. Using a driver cost of US$0.32/hp-hr, this reduction in horsepower would result in a savings of approximately US$7000 per year. However, by examining the data in Table 6, the stage capacity increased from 5.12 to 7.12 MMscfd (0.145 x 10 6 to 0.202 x 10 6 sm 3 /d) after the repair. This is a 39% improvement in flow. The horsepower before was 345 and 365 after only a 5% increase. The goal of analysis is to help operators compress and pump gas at the lowest cost. Breaking this down to cost, and using a driver cost of $0.032/hp-hr, the following conclusion can be made. Before repair: $18,200/MMscf for one year, or to pump 7 MMscfd (0.198 x 10 6 sm 3 /d) for one year, $127,000. After repair: $13,900/MMscf for one year, or to pump 7 MMscfd for one year, $97,000. To pump the same amount of gas through this cylinder, the clogged valves were costing the owner US$30,000 per year. More importantly, the clogged valves on this machine were throttling capacity to 5 MMscfd (0.142 x 10 6 sm 3 /d) whereas the machine should have been capable of 7 MMscfd (0.198 x 10 6 sm 3 /d). This lost capacity resulted in lost revenue for the producer. As previously mentioned, assuming a sale price of US$3.50/MMBtu for the gas, every MMscf of gas is worth US$1.3 million. Therefore, the reduced capacity of 2 MMscfd (0.057 x 10 6 sm 3 /d) of gas for this unit was costing the user US$2.6 million per year in potential revenue! Conclusions Compressor analysis is an invaluable tool that provides direct insight into the health and performance of reciprocating compressors. As part of a condition-based maintenance program, reciprocating analysis has proven to reduce maintenance costs and protect against catastrophic machinery failures. The economic benefit of compressor analysis can be quantified by theoretical models or more directly by examining the driver cost before and after repairs are performed. The fuel savings from a typical discharge valve leak are significant and can quickly provide a return on investment for an analysis program. By improving gas flow through a compressor, an operator can maximize the economic value of a reciprocating compressor and greatly improve revenue generation. CT2 Need Subject Matter Experts for Recip Analysis? WE PROVIDE THE BEST IN THE INDUSTRY. Get the experts you need when and where you need them. Windrock offers the most knowledgeable analysts in the industry. So whether you need assistance with portable analysis, online systems, onsite analysis or recip training you ll gain from our expertise. Se habla español TURNING DATA INTO PERFORMANCE +1 865.330.1100 sales@windrock.com www.windrock.com