Economic Justification of Magnetic Bearings and Mechanical Dry Seals for Centrifugal Compressors

Transcription

1 THE AMERICAN SOCIETY OF MECHANICAL ENGINEERS 345 E. 47 St. New York, N.Y GT-174 The Society shall not be responsible for statements or opinions advanced in papers or in discussion at meetings of the Society or of its Divisions or Sections, or printed in Its publications. Discussion is printed only if the paper is published in an ASME Journal. Papers are available from ASME for fifteen months after the meeting. Printed in USA. Copyright 1987 by ASME Economic Justification of Magnetic Bearings and Mechanical Dry Seals for Centrifugal Compressors S. 0. UPTIGROVE, P. Eng. NOVA, An Alberta Corporation T. A. HARRIS, P. Eng. NOVA, An Alberta Corporation D. 0. HOLZNER NOVA, An Alberta Corporation ABSTRACT The method for calculating the economic justification related to the use of active magnetic bearings and mechanical dry seals in centrifugal compressors is presented. This evaluation is based on the initial purchase price, and the associated operational and maintenance savings over standard oil bearings and oil seal systems. The economic evaluation is presented for both the retrofit of existing equipment and for the installation into new equipment. Although the economics are shown for centrifugal compressors, the same basic philosophy can be used to evaluate the economic justification of magnetic bearings for any type of rotating machinery. Different areas of potential cost savings are discussed, including a method of calculating the associated savings. In addition, an example calculation for a specific compressor is presented. It must be emphasized that the economic benefits and the associated costs are highly dependent on the type and size of compressor as well as the particular application. INTRODUCTION In 1978, NOVA, AN ALBERTA CORPORATION initiated a program to identify, research and implement technologically and economically sound means of eliminating the many problems that were being experienced with the oil systems on its compressors. The two main purposes of this program were to reduce the high maintenance costs and the high number of call outs resulting from these oil system problems. NOVA worked closely for several years with John Crane Houdaille Inc. on the development, application and installation of a mechanical dry seal for centrifugal compressors. By 1980, NOVA had a successful, fully operational mechanical dry seal system in its gas transmission system. Since 1980, NOVA has installed nine mechanical dry seal systems. An additional twelve mechanical dry seal systems are scheduled to be installed in The Active Magnetic Bearing was developed by Societe de Mecanique Magnetique (S2M) of Vernon, France. The first successful application of magnetic bearings on a centrifugal compressor was completed in 1985 and was a co-operative effort between NOVA, S2M, and Magnetic Bearings Inc. (MBI) of Radford, Virginia. NOVA's second magnetic bearing compressor retrofit was completed in 1986 with an additional four installations planned for Together, the magnetic bearings and the dry seal systems have resulted in the development of completely oil-free compressors, which have proven to be a much more economically efficient. NOMENCLATURE Units: SI (U.S. Customary) A = Power required to run the magnetic bearings, W (HP). a = The increase in efficiency of the pipeline due to the elimination of oil contaminants. B = The heating value of the gas, J/m 3 (BTU/SCF). b = Load factor (percentage of total available power used during normal operation). C = Cost per hour of lost production, $/hr. d = The density of the oil, kg/m 3 (lb./gal.) E = Efficiency of the gas generator and power turbine. F = The power rating of the seal oil cooling fan, W (HP). f = A conversion factor for cost per year for power, $/W y ($/HP y) G = The price of the process gas, $/m 3 ($/SCF). H = Lost production time each year due to seal oil system related shutdowns, h/y. h = The hours of operation each year, h/y. Presented at the Gas Turbine Conference and Exhibition, Anaheim, California May 31-June 4, 1987

2 i = Quantity of gas being vented including oil entrapment from the oil seals, m3 /s (SCFM). L 1 = The flow rate of gas through an oil seal, m 3 /s (SCFM). L 2 = The dry seal filtered gas supply flow rate, m 3 /s (SCFM). 1 = Vented gas from the dry seals, m 3 /s (SCFM). M = Average yearly cost for scheduled maintenance on the oil system system, $/y. m = Average yearly cost for scheduled maintenance on the dry seal system, $/y. n = The number of seals in the unit. P = The power required of the driver at normal operating conditions, W (HP). p = The price of the oil being used, $/m 3 ($/BBL). Q = Quantity of oil used each year, m 3 /y (BBL/y). q = The specific heat of the oil, J/kgK (BTU/lb. F). R = The power rating of the seal oil pump, W (HP). S 1 = Savings due to a reduction in power losses in the seals for dry seals vs. oil seals, $/y. S 2 = Savings from the elimination of seal oil pumps and cooling fans, $/y. S 3 = Savings from the reduced gas leakage in dry seals vs. oil seals, $/y. S 4 = Savings due to increased pipeline efficiency, $/Y- S 5 = Savings from the elimination of oil consumption, $/Y. S 6 = Savings due to reduced maintenance and downtime with the installation of dry seals, $/y. S 7 = Savings from the elimination of parasitic oil losses, $/y. S 8 = Savings due to reduced maintenance, call outs and downtime with the installation of magnetic bearings and dry seals, $/y. T 1 = The temperature of the oil going into the cooling system, K ( F). T 2 = The temperature of the oil coming out of the cooling system, K ( F). t = Percentage of call out time and unscheduled maintenance time that are related to the seal oil system. U = Total cost per year for all call outs and unscheduled maintenance, $/y. V = Normal operating flow rate for the compressor, m 3 /s (SCFM). v = The flow rate of oil through the cooling system, m 3 /s (gal./h). W = The total cost of lost production time per year, $/Y. X = The power lost in the oil seals due to friction, W (HP). Y = The power lost in a dry seal, W/mm (HP/in.) Z = The average maintenance cost of the compressor per year, $/y. ECONOMIC JUSTIFICATION OF MECHANICAL DRY SEALS Oil seals, including their auxiliary equipment such as pumps, drainers, back pressure valves, relief valves, degasser tanks, seal oil coolers, fans, etc., can all be eliminated with the use of mechanical dry seals. The cost of a standard seal oil system is generally more than or equal to the cost of a mechanical dry seal system. The actual cost difference will depend on the size of the equipment. There are also substantial cost savings that can be realized from improved efficiency and reduced maintenance and operating costs. The efficiency is increased by eliminating parasitic power losses due to auxiliary pumps and oil shear. Depending on the unit size, there are potential savings of up to 150 kw (200 HP). Major savings are also possible in the operation of the pipeline. Seal oil leakage into the pipeline has an adverse effect on pipeline efficiency. The oil can also affect meter readings. In process applications, oil often causes product contamination requiring process interruption. Any of these problems can necessitate cleaning the pipe and equipment, or removal of severely contaminated equipment. With a dry seal system these problems are eliminated. To do an economic evaluation of the use of dry seals versus oil seals in existing or new compressors, the costs associated with operating and maintaining both types of systems must first be evaluated. These costs will vary considerably depending on location conditions and the size and type of compressor. However, it is possible to identify and evaluate the following key cost areas: (a) The power loss in shaft rotation for oil seals versus dry seals. (b) The power used by the seal oil pump and cooling fans. (c) The power used to recompress gas used in the seals or the cost of that gas if vented to atmosphere from the seals. (d) The increase in pipeline flow efficiency due to the removal of oil leakage. (e) The oil lost through the oil seals. (f) The maintenance, call outs and downtime required on an oil seal system versus a dry seal system. In addition to these key areas of cost savings, there is the increased safety aspects of dry seals from the elimination of the high pressure oil system. This increased safety is very difficult to put a dollar value on but in many cases is the main reason for a retrofit. In order to evaluate the above operating and maintenance costs, specific information must be obtained about the seal oil system. This includes the power loss through oil shear in the seals, which can be obtained from the compressor or seal manufacturer. It can also be calculated by measuring the temperature of the oil going to and coming from the seals. By taking the temperature rise in the oil across the seals together with the flow rate and the specific heat of the oil, the oil shear losses can be calculated as shown in Eq. (8). This assumes all the power lost through oil shear in the seals is converted to heat. Some approximation may be necessary if the seal and bearing oil drains are combined. Additional information that is required regarding the seal oil system includes the quantity of gas being vented to atmosphere from the degassing chamber and the quantity of gas fed to the wet seals from the - 2 -

3 discharge. These values can also be obtained from the manufacturer or measured on the existing seal oil system. Given the required information about the seal oil system and the price of gas and its heating value, the costs associated with each of the six key areas of savings can be calculated as follows: (a) The power loss in shaft rotation from oil seals versus dry seals. In order to simplify these calculations, first calculate the conversion factor for dollars per year per unit of power for the compressor in question. This is done as follows for a gas turbine driven unit: f _ Gxhxb B x E Similar calculations can be done to establish a price per year per unit of power for motor drives or steam turbine drives. The next step is to calculate the savings per year for dry seals versus oil seals due to power losses in the seals: S 1 = n (X - Y) f (2) The power lost in a dry seal is approximately 5.87 W/mm (9.2 HP/in.) of sealing diameter [i.e., a mm (6*") sealing diameter seal would use 0.9 kw (1.25 HP)]. (1) (c) The power used to recompress gas used in the seals or the cost of that gas if vented to atmosphere from the seals. In many oil seals, buffer gas from the compressor discharge is injected into the seal to try to prevent oil from entering the process gas. With dry seals, discharge gas is filtered and injected into the seals before the sealing faces. Both of these practices result in a recycling of gas from discharge to suction that wastes power. This power loss is taken into account in the first part of Eq. (4). The amount of gas escaping from the oil seals can be calculated by measuring the volume and percentage of gas being vented from the degassing tank. Alternately, this figure may be obtained from the manufacturer. If, the gas escaping from the wet seal is returned to compressor suction, the only gas escaping is through oil entrapment. The amount of gas lost in this manner is approximately the same as the dry seal leakage. However, if the seal oil trap is vented to atmosphere instead of back to the suction in order to stop product oil contamination, the leakage can be as much as 100 times that of a dry seal. The cost of gas being vented to atmosphere is taken into account in the second part of Eq. (4). The quantity of gas being vented from a dry seal is approximately 5.57 x 10-6 m3 /s per mm of sealing diameter (0.3 SCFM/inch of sealing diameter). (b) The power used by the seal oil pumps and cooling fans. The power ratings of the pumps and fans are generally shown in the compressor manual. If the pumps are mechanically driven by the compressor unit, the power consumption at full speed should be used as the conversion factor "s" takes into account the power rating versus the normal operating conditions. The pumps are sized to provide adequate flow at minimum operating speed. At higher speeds the extra oil pumped is merely diverted back into the oil tank. There are a great many arrangements for an oil system and compressor. As a result, care must be taken when evaluating a retrofit that the oil no longer required by the seal can actually be saved. An example of this occurs when the seal oil pump also drives the cooling fans for a common seal and bearing oil cooler. The reduced flow caused by a dry seal will not actually result in a power saving unless the pump is resized for the lower flows. If the pump is not resized the extra flow is returned to the oil tank. On a new unit this power is saved as the pumps would be properly sized. Using the power rating for the seal oil pumps and fans, the savings per year from the elimination of this equipment can be calculated as follows: S 2 = (R + F) f (3) (d) S 3 = [n(l 1 - L 2 ) x 1,; x f] + [n(i - 1) x G x h] Increased pipeline flow efficiency due to the removal of oil leakage. Except for oil loss in oil spills within the compressor building, all lost oil enters the piping. This results in the pipeline being coated with oil as well as oil collecting in low points within the pipeline system. This oil will collect solid particles in the gas stream and continue to build up. All these factors increase the effective roughness of the pipe while reducing the volume. Many studies have been done on the effects of contaminants in the pipeline such as oil, condensates and rust. The effects of these contaminants on the efficiency of the pipeline depends on many parameters such as the size of the pipeline, whether flow is turbulent, and the terrain or number of low areas. Although these contaminants can reduce the overall volume of the pipe the significant effect is the increased effective roughness in the pipeline which is not a problem unless the gas flow is turbulent. This will occur when the pipeline is running at its capacity. Test results have shown [2] that after 10 years of service these contaminants can reduce the flow by as much as 5% which can be regained by cleaning or pigging the line. In many cases where the process gas is clean, dry natural gas, the only source of contaminants is the oil from the compressor which - 3 -

4 then collects solid particles. On the other hand, if the gas contains significant amounts of liquids the elimination of the seal oil leakage into the pipeline will not eliminate this contaminant build up. Additional information on determining pipeline efficiency and the effects of contaminants can be found in References [1] and [2]. An increase in efficiency would not be noticed immediately after a dry seal installation unless the pipeline is cleaned and no other seal oil system is leaking into the pipeline. The savings from this increase in flow efficiency would equate to an equal decrease in the power required to compress the same flow rate which can be written as follows: S4 = a x P x f (5) (e) The oil lost through the oil seals. The elimination of oil consumption can be a key area of cost savings depending on the type of oil being used and the amount being lost. The total oil consumption per year should be established from actual operating records both from normal seal leakage and from seal and seal oil system failures. Often the process compressor is required to stay on-line even when the seal oil consumption is well above normal. The cost of the oil will vary with the type used and the associated transportation costs. S5 = Q x p (6) (f) The maintenance, call outs and downtime required on an oil seal system versus a dry seal system. The high cost of maintenance on oil seal systems and the related downtime is the main reason that development and application of mechanical dry seal technology was started. There is very little that can go wrong with a mechanical dry seal system, as the control panel is nothing more than a safety monitoring device which monitors and filters the process gas flow to the seal and monitors the vent gas from the seal. In the seal itself, there is no contact between the moving and stationary parts during operation, and thus there is virtually no wear. An oil seal system on the other hand, has pumps, fans, and the seals themselves which require scheduled maintenance and often unscheduled maintenance due to parts wearing out. The dry seal monitoring system is much less complicated than a seal oil system. In addition, seal oil systems vary from one manufacturer to the next. This requires maintenance personnel to be trained and familiar with each type of compressor. The costs associated with this maintenance is taken into account in the middle part of Eq. (7). Operator call outs to unmanned stations due to seal oil related problems is another key area of cost savings. Within the NOVA Gas Transmission System an average of 50% of all call outs to the stations are due to problems related to the seal oil system. Where dry seals have been installed, call outs due to the seal system have been virtually eliminated. The reduced number of call outs and the associated costs are taken into account in the first part of Eq. (7). Finally, if there is no backup compressor system, the additional cost of lost production during scheduled and unscheduled downtime due to seal oil related problems should also be taken into account which is the last part of Eq. (7). 5 6 = (t x LI) + (M - m) + (H x C) (7 ) If there is a backup compressor, the cost of having this additional equipment available should be included in the evaluation. By comparing the cost of retrofitting a compressor with mechanical dry seals to the sum of the above six areas of cost savings, the associated payback period can be calculated. The following sample economic evaluation demonstrates the use of these equations and the economic benefits of the mechanical dry seal. Sample Economic Evaluation The following sample economic evaluation is based on an actual unit operating in the NOVA gas transmission system and is shown in Canadian dollars. The compressor, with a 165 mm (6.5 inch) diameter shaft and a 190 mm (7.5 inch) sealing diameter, has two seals. This unit runs 8000 hours each year and has an overall gas generator and power turbine efficiency of 25%. Each seal uses 35 kw (47 HP) in rotating oil shear losses. The seal oil pump is 19.4 kw (26 HP) and the cooling fan is 7.5 kw (10 HP). The heating value of the gas is MJ/m 3 (1000 BTU/SCE) and costs $88.7/10 3m3 ($2.50/MSCF). The seal buffer gas is taken from the balance piston instead of the compressor discharge and the seal oil traps return the seal gas to suction. The oil seal gas leakage is 3.0 SCFM per seal due to oil entrapment. There is approximately 0.1 m 3 (0.6 BBL) of synthetic oil added to the reservoir each week at a cost of $912/m 3 ($145/BBL). The turbine is site rated at 8700 kw (11,650 HP) with a compressor flow rate of 26 m 3 /s (690,000 SCFM) and normally operates at full load. The unit has a backup system so there is no lost production during downtime. There is an average of $7,000 spent on seal oil system maintenance for the unit each year. The dry seal system has an average yearly maintenance of $2,000. The percentage of call outs and unscheduled maintenance related to seal oil system problems is about 50% with the total cost of all call outs and unscheduled maintenance for the unit running around $30,000/year. (a) The savings due to the power lost in oil seals vs. dry seals: - 4 -

5 f _ Gxhxb B x E (1) (f) Cost savings due to reduced maintenance, call outs and lost production time. f - $88.7/10 3 m 3 x 8000 h/y x 3600 s/h x x 10 6 J/m 3 x = (t x U) + (M - m) + (H x C) (7) f = $0.27/W x year S I = n (X - Y) f (2) S 1 = 2 (35,000 W W/mm x 190 mm) x $0.27/W x year 5 1 = $18,298/year (b) The savings from the elimination of the seal oil pumps and cooling fans: (In this case the cooling fan horsepower could not be saved without replacing the hydraulic fans as the seal and lube oil runs through a common cooling system.) S 2 = (R + F) f (3) = (19,400 W + 0) x $0.27/W x year S 2 = $5,238/year (c) The costs associated with the power used to recompress gas used in the seals and the cost of the gas vented to atmosphere from the seals. In this case the gas entrapment in seal oil was estimated to be 1.42 x 10-3 m3 /s (3 SCFM) and the gas used for the oil seals was taken from the balance piston instead of the discharge, thus the amount of gas recompressing is taken to be zero. The filtered gas supplied to the dry seals is 9.44 x 10-3 m 3 /s (10 SCFM). S 3 = [n(l 1 - L 2 ) x x f] + [n(i - 1) x G x h] (4) k W = [2 ( x 10-3 m 3 /s) x 8700 x $0.27/W x year] 326 m 3 /s + [2 (1.42 x 10-3 m3 /s x 10-6 m3 /s x 190 mm) x $88.7/10 3 m3 x 8000 h/year x 3600 s/l] = -$68/year + 1,848/year 53 = $1,780/year (d) Cost savings due to the increase in pipeline flow efficiency: (In this case the process gas is dry natural gas operating in the turbulent flow range and the increase in pipeline efficiency due to the elimination of oil in the line is taken to be 1.5%. This is estimated from the information in Reference [2] which shows the effect of contaminants in a pipeline can effect the efficiency of the pipeline by as much as 5%). 54 = axpxf = x 8,700 kw x $0.27/W x year 5 4 = $35,235/year (e) Cost savings due to elimination of oil losses: S 5 = Oxp = 0.1 m 3 /week x 52 weeks/year x $912/m 3 S 5 = $4,742/year (5) (6) In this case H = 0 as the backup system prevents lost production. = (0.5 x $30,000) + ($7,000 - $2,000) + (0) S 6 = $20,000/year TABLE 1 - SUMMARY OF COST SAVINGS FOR DRY SEALS AREA OF COST SAVING SAVINGS/YEAR (a) parasitic seal losses (5 1 ) $18,298 (b) elimination of seal oil pump (5 2 ) $ 5,238 (c) recycled gas and gas losses (5 3 ) $ 1,780 (d) increase flow efficiency (5 4 ) $35,235 (e) oil consumption (55 ) $ 4,742 (f) maintenance, call outs and downtime (5 6 ) $20,000 Total Savings $85,293/year Dry Seal Installation and Retrofit Costs The cost of dry seals will vary with the size of the machine and whether it is a beam or overhung unit. The cost of the seals varies directly with the size of the sealing diameter. The sealing diameter is usually approximately one inch larger than the shaft diameter. The costs are reduced for the retrofit of an overhung compressor as only one seal is required. With a beam type of compressor, two seals are required. A major criteria for a successful seal retrofit is whether the seal will fit into the compressor with only minor modifications to the shaft and housing. If a new shaft and housing are required, the cost of a retrofit could be doubled. To date, all seal retrofits have been accomplished without requiring new shafts or housings. The cost of the dry seals can also be substantially reduced if several identical machines are retrofitted at the same time. This is because the costs can be spread over several seals. As more retrofits are completed, the seal manufacturers will develop standard seals for most of the different types of machines, thus eliminating the prototype design cost. The dry seal has minimal effect on the rotordynamics of large machines, but the rotordynamics must be examined on small, high speed compressors as the weight of the seal becomes high in relation to the shaft. In any event, the rotordynamics should be checked for every installation. Dry seal retrofits require minimal station modifications. The controls for the dry seal can usually use the same wires and annunciation that are being used by the existing oil seal system. The cost of controls and piping to the unit is generally constant regardless of unit size as the same instrumentation is used. The installation of a dry seal system in a new compressor will generally cost less than a conventional seal oil system. With the substantial - 5 -

6 cost savings in operation and maintenance, the use of wet seal oil systems may ultimately become obsolete. TABLE 2 - SUMMARY OF INSTALLATION COSTS AND THE ASSOCIATED OPERATING AND MAINTENANCE COSTS (FROM EXAMPLE) DRY SEAL OIL SEAL Installation in a New Unit $ 90,000 $100,000 Retrofit of an Existing Unit $180,000 Average Yearly Operation Cost $ 5,000 $ 90,000 Yearly Savings $ 85,000 seals. In this case the temperature difference across the oil cooling system can be used to calculate the total parasitic power losses in the bearing and seal oil system. This calculation is based on the assumption that all horsepower used by the seals, pumps, bearings and oil shear in the piping is converted to heat which must be transferred out of the oil by the cooler. The value calculated from this equation is generally a little low as there is still heat being dissipated through the piping and in the oil tank. Payback for Retrofit 2.1 years S 7 = [(T1 - T2)dxqxv- A]xf (8) It must be emphasized that these figures came from an economic evaluation for a specific unit and that the costs associated with each unit will depend on the type of unit and its application. In addition, the payback would be further reduced by taking into account the net present value after tax. This has not been done in this case in order to simplify the evaluation. ECONOMIC JUSTIFICATION OF ACTIVE MAGNETIC BEARINGS AND MECHANICAL DRY SEALS When compared with the conventional lubricated bearing system, the installation of magnetic bearings in new compressor equipment will generally cost slightly more. Depending on the size of the equipment, the additional cost is usually between 10% and 50% of the price of the unit. This additional up-front cost is offset by the significant savings in operating and maintenance costs. By eliminating parasitic horsepower losses and oil shear, as much as 2% of the unit's output power can be saved. The magnetic bearing system on a compressor uses less than 4 kilowatts of energy, compared to the 225 kw (302 HP) lost in the conventional bearing and seal oil system on the retrofit of the compressor shown in the example. The subsequent retrofit of the power turbine on this unit can net an additional saving of 325 kw (436 HP) with the elimination of oil shear, pumps and cooling fans. In addition, magnetic bearings eliminate mechanical wear allowing unlimited bearing life, and reduce maintenance costs leading to an overall improvement in system integrity and reliability. It was the successful development of the dry seal system which made the application of magnetic bearings to a compressor a viable alternative. It would have been pointless to eliminate the lubrication system and keep the seal oil system with its associated problems and the possibility of the seal oil affecting the magnetic bearings. It is for this reason the economic justification for magnetic bearings should include both the costs and savings associated with a mechanical dry seal system. The economic evaluation for the use of magnetic bearings and mechanical dry seals in existing equipment and in new equipment can be carried out in much the same fashion as the economics of the dry The other cost savings are calculated in the same way as shown in the economic evaluation of dry seals with the exception of the maintenance and reliability, Eq. (7). The installation of magnetic bearings in conjunction with mechanical dry seals results in a totally oil-free compressor. Virtually wear-free seals and bearings eliminate the requirement for any auxiliary oil equipment as well as substantially reduce machine vibration. For the case shown in the example the overall maintenance, call outs and downtime were reduced by 85%. Thus, Eq. (7) becomes: S 8 = 0.85 (Z + U + W) (9 ) Sample Economic Evaluation Using the same compressor as was used for the sample economic evaluation in the dry seal section, the following additional cost savings can be obtained through the use of magnetic bearings. The temperatures of the oil going in and out of the cooling system are 337 k (147 F) and 322 K (120 F) respectively. The oil density is kg/m3 (7.095 lb. mass/gal.) with a specific heat of 2177 J/kgK (0.52 BTU/lb. F). The flow rate of oil through the cooler is 8.45 x 10-3m3 /s (134 U.S. gpm). The total average maintenance cost for the unit and associated equipment is $55,000. S 7 = [(Ti - T2)dxqxv- A]xf (8) = [( )K x kg/m3 x 2177 J/kgK x 8.45 x 10-3 m 3 /s W] x $0.27/W x year S 7 = $62,262/year The savings due to the required maintenance call outs and downtime would also change to: S 8 = 0.85 (Z + U + W) (9) = 0.85 ($55,000 + $30, ) S 8 = $72,250/year The following table summarizes the savings per year for the installation of both magnetic bearings and dry seals.

7 AREA OF COST SAVINGS TABLE 3 - SUMMARY OF SAVINGS FOR MAGNETIC BEARINGS AND DRY SEALS Parasitic oil system losses (S7 ) Recycled gas and gas losses (5 5 ) Increased flow efficiency (5 4 ) Oil consumption (S 5 ) Maintenance, call outs and downtime (S5 ) Total Savings SAVINGS/YEAR $ 62,262 $ 1,780 $ 35,235 $ 4,742 $ 72,250 $176,269/year The following table summarizes the installation costs, the operating and maintenance savings and the resulting payback for the specific compressor used in the example. TABLE 4 - DRY SEALS AND MAGNETIC BEARINGS COSTS AND SAVINGS Additional cost in new unit $150,000 Retrofit of an existing unit $775,000 Yearly savings $176,269 Payback for a new unit Payback for the retrofit less than 1 year 4.4 years As shown in this sample evaluation, significant reductions in the cost of operating compressor stations are possible with payback in as little as four (4) years. This includes the cost of retrofitting both the oil seal system with mechanical dry seals and the lube oil bearing system with magnetic bearings. Once again, this does not take into account the net present value after tax which would further reduce the payback period. Additional savings can also be seen in overall compressor building size requirements. If new installations are designed with magnetic bearings and mechanical dry seals, as much as 50% of the space required can be reduced with the elimination of oil tanks, coolers and other oil system equipment. At locations where space is at a premium, this reduction in space can be very important. A significant decrease in the total package weight will also be accomplished. Another advantage is the magnetic bearing's monitoring system which can be used as a diagnostic tool for establishing and monitoring the operating parameters of the equipment. This makes it possible to determine the rotor imbalance during operation and easily rebalance the equipment at operating speeds. Improvement in the rotordynamic stability of the equipment will occur and vibration will be substantially reduced by operating the equipment within its inertial axis as opposed to its geometric axis. This automatic balancing feature of the magnetic bearings will also reduce internal labyrinth seal wear. Internal recirculation losses can be from 0.5 to 5.0%, depending on the compressor size and flow conditions as well as the wear in the internal labyrinth seal. This increase in losses due to wear can be very costly. For example, an increase in balance seal loss from the design of 1% to 2% in worn condition would cost $23,490 per year for the compressor shown in the example. This loss is not reduced by installing magnetic bearings unless the labyrinth seal is replaced at the same time but the magnetic bearings will prevent further wear. All of this contributes to a more reliable, maintenance free system that will increase the economic benefits to the consumer. As has been stated, the number of areas of economic benefit are dependent on each application. In each case, however, the economic benefits and the technical advantages are significant. CONCLUSION Active magnetic bearings in conjunction with mechanical dry seals make it possible to operate centrifugal pipeline compressors without seal or lube oil systems. Mechanical dry seal systems have been shown to be both technically and economically sound, on both new equipment and as a retrofit to existing equipment. The retrofit of existing equipment with both these technologies does have to be evaluated from an economic standpoint on a site-specific basis. In most cases, such a retrofit program has shown a payback of between three (3) to six (6) years. In addition to these economic benefits, there is the increased safety through the elimination of the high pressure seal oil system and the lubrication system. By eliminating the requirement for oil you also eliminate the chance of oil related fires. Due to these economic and safety benefits of mechanical dry seal and magnetic bearing systems, companies have started to retrofit their existing machines as well as specifying this equipment in new machines. Some of the retrofits scheduled for 1987 include: a retrofit of a 9,556 kw (12,809 HP) Ingersoll-Rand power turbine with magnetic bearings, a 8,579 kw (11,500 HP) Cooper Bessemer compressor, a 2,641 kw (3,540 HP) Solar compressor, and a 3,823 kw (5,125 HP) Delaval compressor, and a 2,633 kw (3,530 HP) Ingersoll-Rand compressor all retrofit with both magnetic bearings and dry seals. New installations with magnetic bearings and dry seals scheduled for 1987, include an Ingersoll-Rand CBF-842RE compressor and a Delaval 3B-37 compressor. In addition, many mechanical dry seal system retrofits are being carried out in both sweet natural gas and toxic gas applications as well as process applications. REFERENCES [1] Propan, V. A. "Pipeline Efficiency Testing Part I: Theory and Method." Pipeline and Gas Journal 206/2 (February 1979): [2] Propan, V. A. "Pipeline Efficiency Testing Part II: Measurements and Calculations." Pipeline and Gas Journal 206/6 (May 1979):

8 [3] Foster, E. G., V. Kulle and R. A. Peterson. "The Application of Active Magnetic Bearings to a Natural Gas Pipeline Compressor." ASME Paper 86-GT-61, [4] Hesje, R. C. and R. A. Peterson. "Mechanical Dry Seal Applied to Pipeline (Natural Gas) Centrifugal Compressors." ASME Paper 86-GT-3, [5] Habermann, H.and M. Brunet. "The Active Magnetic Bearing Enables Optimum Control of Machine Vibrations." ASME Paper 86-GT-221, [6] Weise, D. A. "Active Magnetic Bearings and Their Industrial Applications." Magnetic Bearings Inc., June 1985.