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MITIGATION OF SEASONAL PRODUCTION LOSS FOR THREE PARALLEL 4.7 MMTPA LNG TRAINS

Nicholas White

Machinery & Reliability Senior Advisor RasGas Company Ras Laffan, Qatar

Divyesh Master Head Onshore Development

RasGas Company Doha, Qatar

KEYWORDS: Compressor Restage, Gas Turbine Inlet Cooling, Production Loss, Seasonality Mitigation ABSTRACT

RasGas’ production facility at Ras Laffan, Qatar has seven LNG trains which were installed and commissioned between 1999 and 2010. Trains 3, 4 and 5 each have a nameplate production capacity of 4.7 MMTPA and the latter two produce lean LNG. The trains are of the Air Products Inc. (APCI) propane, mixed refrigerant design. Each refrigeration compressor string is driven by a General Electric (GE) Frame 7EA gas turbine and an ABB starter/helper electric motor. During the hot summer months, production reduces by as much as 5% resulting from reduced gas turbine power and greater cooling water temperature. An optimization initiative was therefore, conducted to remove this seasonality impact. The project consisted of three main aspects:

• revision of process conditions for the refrigeration cycle which incorporated the raised production capacity

• design for more gas turbine power on the LP/MP-MR compressor string through gas turbine inlet air cooling

• design a restaged propane compressor with satisfactory operating envelopes for the revised process conditions

This paper focuses upon the adopted methodology which successfully led to the consolidated conceptual design. It also provides insight to the project management challenges where multiple parties were contributing to common goals. Finally, the proposed execution schedule and project investment is discussed.

INTRODUCTION

RasGas operates LNG Trains 1 to 7 with an annual LNG production of 37 MTPA and a sales gas plant supplying 2 BCFD of treated gas. Trains 1 to 3 are designed for rich LNG production, Trains 4/5 are designed for rich and lean LNG production and Trains 6/7 for lean LNG production. These facilities are located at Ras Laffan Industrial city, in the north of Qatar. Due to normal changes in the ambient conditions, resulting in increased air and sea water temperatures, the LNG production capacity of Trains 3 to 5 reduces by approximately 5% (termed as seasonality impact) during the summer months. This paper discusses a feasibility study which was led by the Venture Planning Department to remove seasonality impact on LNG production for Trains 3, 4 & 5.

A study team was formed comprising of engineers from RasGas (Company), Elliott Group (Compressor OEM), General Electric (GT OEM) and APCI (Refrigeration Licensor). Chiyoda Corporation (EPC Contractor) was selected as the engineering contractor to lead and co-ordinate the study on behalf of RasGas since it was the appointed EPC contractor during design of the LNG trains. Figure 1 describes the organization and working structure of the feasibility study team.

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Figure 1. Organization Chart for Study Team

A basic process flow diagram of Trains 3, 4 and 5 is illustrated in Figure 2 (see Appendix). The refrigeration cycle is the typical APCI C3-MR design (Roberts, et al. 2002). The rotating machinery configuration is also shown. There are two compressor strings, the LP/MP-MR string and the HP-MR/PR string. In both strings, the compressors are connected together by large flexible couplings. They are also driven at opposing ends, again through flexible couplings, by a GE Frame 7EA gas turbine and ABB 16086 HP (12 MW) helper motor at 3600 rpm rated shaft speed.

PRODUCTION LOSS DUE TO SEASONALITY

Trains 3, 4 & 5 were originally designed in 2002 and then commissioned in 2004, 2005 and 2006 respectively. Since design and commissioning, several changes in the ambient conditions (air temperature, relative humidity, sea water temperature) have been observed, which directly impacts the LNG production capacity. Another observation was a general increase in gas turbine (GT) intake air temperature after construction of Trains 6/7 and sales gas facility AKG2 in 2010 & 2011. Hence, prior to initiating detailed studies amongst the parties of Figure 1, it was decided to revalidate the ambient conditions which impact LNG production. A detailed description of how this was done is given in a later section but it should be understood here, that it was a comprehensive task involving approximately 82,000 meteorological data points.

GT Hot Air Recirculation

Trains 6, 7 and AKG2 were constructed adjacent to Trains 3, 4, 5 and AKG1. The plot plan for AKG2 is segregated to fit the available plot space. Recent observations indicated slightly higher than normal GT inlet air temperatures. To verify this, the project team engaged the EPC Contractor to review whether hot air recirculation from the adjacent facilities was contributing to elevated GT air intake temperatures. It should be noted, that a 1.0 deg C increase in GT air intake reduces the Frame 7EA GT power by approximately 0.5%.

The review was based on historical ambient data taken over 10 years (2001 – 2010) and included prevailing wind directions. It was confirmed that hot recirculation was occurring at site mainly from the fin-fan air coolers. Various recommendations were provided to mitigate the impact on reduced GT output. For example, increasing the stack heights of the various hot air sources. Unfortunately, the majority were impractical for implementation due to cost, risk of implementing on an operating plant and the required downtime to complete the modifications. As such, to allow for this impact, the project design basis was updated by an additional 1.5 deg C over the average ambient temperature. Figure 3 (see Appendix) shows a plot plan of the whole facility with the highlighted areas indicating sources of hot air.

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Reduced Stream Day Capacity

The stream day capacity throughout the year is shown in Figure 4 (below) for Trains 3 and 4. Between April and November both trains begin to lose production capacity when compared to the winter capacity.

The EPC Contractor and Helper Motor OEM jointly completed a study to understand the feasibility of upgrading the helper motor capacity. It was concluded that, although extra capacity would be possible without impacting the motor footprint, significant modifications to the underground power supply cable would be required. This was deemed impractical due to the risk associated with carrying out such modifications within an operating facility. Power augmentation by upgrading the helper motor was therefore, not selected.

Figure 4. Reduced Train 3/4 SDC During Summer Months

Following the SDC reduction assessment, the focus areas listed below in Table 1 were selected to remove the summer seasonality impact.

Table 1. SDC Reduction Focus Areas

Cooling water

EPC Contractor

Additional heat duty during summer

Compression Compressor OEM

Assess capacity of MR and PR compressor

Hot air recirculation

EPC Contractor

Allow for elevated temperature to calculate available power from the GT units

Power GT OEM GT Inlet air cooling Cryogenics Refrigeration

Licensor Review capacity of the MCHE and other exchangers in the liquefaction circuit

REFRIGERATION LICENSOR STUDY

An assessment of the liquefaction facilities for Train 3 (Rich LNG) and Train 4/5 (Lean LNG) was carried out by the Refrigeration Licensor. Identification of any bottlenecks in the liquefaction system was paramount. A Base Model was developed for the simulation to trace the actual performance of liquefaction, refrigeration, and fractionation units based on 2010 August operating data. The EPC Contractor supported the study by providing the required design data (heat exchangers, vessels etc.) for developing the model. This step was necessary to re-validate the Refrigeration Licensor’s model, which was developed for detailed design using current operating data. Key observations from the Base Models for Train 3 and Train 4/5 were as follows:

1. The primary limitation to LNG production was the power available to the LP/MP-MR compressor string during summer period

2. There was unused power available to the HP-MR/PR compressor string in summer (and winter).

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3. The propane system had adequate capacity to provide pre-cooling in summer.

The conclusion drawn from the Base Models was that in order to increase summer production for Trains 3, 4 and 5 an increase of available power to the LP/MP-MR compressor string would be necessary through GT inlet air cooling. GT power limitation during the summer was the primary restriction on production. Re-Rating Case simulation models were then created, derived from the Base Models. They provided the necessary modifications to increase the summer rate of production of these trains.

Re-Rating Case 1 – Trains 3 and 4

For Re-Rating Case 1, additional power was made available to the LP/MP-MR compressor string in Trains 3 and 4 by adding chiller coils to the GT air intake. Furthermore, the FCW system was allowed to absorb up to 10% additional duty compared to the Base Case Models, if required. The simulation models of Train 3 and Train 4 were used to rate the performance of the equipment in the liquefaction and refrigeration units, and predict the LNG production increase resulting from the additional available power and higher cooling water duty. Key observations from the Re-Rating Case 1 results for Train 3 and Train 4/5 were as follows:

1. After the addition of power to the LP/MP-MR compressor string by providing inlet air cooling, the primary limitation to LNG production was the operating point of the HHP (4th) stage of the propane compressor which reached stonewall

2. There was still unused power available to the HP-MR/PR compressor string in both summer and winter

3. No constraints were observed on the propane system for pre-cooling

4. The allowed 10% increase in the fresh cooling water duty was only partially consumed

The conclusion drawn from the Re-Rating Case 1 models was to propose a modification of the propane compressor to move its HHP (4th) stage operating point away from stonewall.

An iterative study was carried out between the Compressor OEM and Refrigeration Licensor to produce revised compressor performance maps which could meet the re-rating intent. The base line maps were validated upon shop test data, taken during the original factory acceptance tests. Maps of all four compressor stages were included in the study, since adjusting the duty in one stage had an influence in the other stages for fixed suction and discharge conditions. It was a challenging task because of the applied hardware and aerodynamic constraints and is discussed in detail in a later section.

The simulation model of Train 4 was used to rate the performance of the equipment in the liquefaction and refrigeration units. Increased LNG production was also predicted, resulting from the additional available power, modified propane compressor and increased FCW duty. The Train 4 simulations were performed at both winter and summer conditions. A 7% minimum overload (stonewall) margin was prescribed for the propane compressor stages.

Re-Rating Case 2 - (Summer – Train 4) In summer (Re-Rating Case 2), the inlet air cooling system on the LP/MP-MR GT was in operation and supplied 9.4 MW (11%) additional power to the compressors for the 3200 USRT duty. The FCW duty was 5.4% more than the duty for the Train 4 Summer Base Case Model. The HHP (4th) stage of the Propane Compressor was operating within the allowable margin from stonewall. The installed power available to the HP-MR/PR compressors was fully consumed. LNG rundown was predicted to be increased sufficiently to fully remove the seasonality impact.

Re-Rating Case 3 – (Winter - Train 4)

In winter, the inlet air cooling system on the LP/MP-MR GT was not in operation. The currently installed power available to the LP/MP-MR compressor string was fully consumed. The FCW duty was less than the duty for the Train 4 Summer Base Case Model. The HHP (4th) stage of the Propane compressor was operating within the allowable margin from stonewall. LNG rundown was maintained as per original design basis.

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The Refrigeration Licensor also undertook a study to optimize the capacity of the GT inlet air cooling unit. A sensitivity study was conducted to assess the optimum cooling unit (640 USRT, 800 USRT up to 3200 USRT). The study concluded that the cooling unit with 800 USRT capacity was sufficient to remove the seasonality impact. Table 2 (see Appendix) summarizes the various cases analyzed by the refrigeration licensor to assess the power required for removing the seasonality impact. A sensitivity analysis was performed for Summer; Summer (max) with higher IAC capacity maximizing propane compressor capacity; Winter without IAC; Winter (max) with IAC; and Extreme Summer for highest summer temperature to assess operability.

Re-Rating Case 4 - Train 3

Train 3 produces rich LNG, and hence the overall feed gas rate is approximately 15% less than Trains 4 and 5 (Lean LNG). To optimize the project cost, it was decided to study whether seasonality impact on Train 3 can be removed by just installing LP/MP-MR GT inlet air cooling with no modifications to the propane compressor. Overall, the Train 3 basis is the same as for Train 4 Re-Rating Cases with the following additions or exceptions.

• If the minimum monthly LNG production targets could be met or exceeded without modifications to the propane compressor, Train 3 simulations with the existing propane compressor would be performed.

• The Compressor OEM provided propane compressor performance maps for all stages validated on shop test data.

The study confirmed that seasonality impact on Train 3 LNG production can be removed by installing inlet air cooling unit on the LP/MP-MR GT and operating the propane compressor at reduced speed to maintain a 7% stage overload margin. In addition to that, the FCW supply temperature during winter must be warmed up by isolating some of the FCW/SW heat exchangers. Table 3 (see Appendix) summarizes the various cases analyzed by the Refrigeration Licensor to assess the power required for removing the seasonality impact without changing the propane compressor but with IAC on the LP/MP-MR GT. A sensitivity analysis was performed for summer and winter conditions for various operating speeds.

COOLING WATER STUDY

Assessment of the required FCW conditions to remove the seasonality impact was led by the EPC Contractor. Each LNG train utilizes a closed loop FCW circuit to remove the heat duty from the propane and mixed refrigerant compressors as depicted in Figure 5. Heat duty from the FCW circuit is removed by sea water (SW) by plate-type heat exchangers.

PR Turbine/Compressor MR Turbine/Compressor

Feed Gas

Cooling water/Sea water circuit

Figure 5.Simple Heat Sink Arrangement

The objective of this study was to assess the FCW network and SW/FCW heat exchanger, to specify the FCW and SW supply temperatures after seasonality impact is removed from Train 3, 4 and 5. Temperatures calculated from this analysis were applied in the assessment of the capacity of the liquefaction unit. The FCW hydraulic model of Trains 3 and 4 were developed by the EPC Contractor using in-house software “HYDRONET 2.0”.

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The study classified the FCW heat exchangers as “major” or “minor” based on FCW distribution through each unit as a percentage of FCW flow rate through the circuit. Major heat exchangers were defined as units with FCW flow more than 1% of the total flow rate (approximately 50 TPH) through the circuit. Minor heat exchangers were defined as units with FCW flow below 1% of the total flow rate through the circuit. Detailed performance analysis of the minor heat exchangers was not performed. Simulation and analysis of FCW heat exchangers was based on SW operating temperature data recorded between 2007 and 2010.

For heat exchangers where temperature transmitters were not available, local temperature gauges were temporarily installed to assess the performance. Simulation runs were carried out for winter and summer operating conditions and revalidated against actual plant operating data. FCW and SW operating data were analysed on a monthly average basis in line with evaluation of the LNG production rate and its corresponding sweet gas rate on a monthly average basis. Figure 6 (below) is a sketch for a typical analysis.

SW In

SW Out

SW

/FC

W

Exchanger

All CW users

FCW Pump

FCW Out

FCW IN

Figure 6. Typical FCW/SW circuit

After this analysis, the expected heat duties for a centralised GT inlet air cooler package (one chiller package for three GT units) and a localised GT inlet air cooler package (one chiller package per GT unit) were added and the results are summarized in Table 4 (below).

Table 4. FCW/SW Analysis

Seasonality Removal including Centralised GT

Inlet Air Cooling Package Seasonality Removal including

Localised GT Inlet Air Cooling Package

Flow Rate

[Ton/h] Supply Temp

[°C] Return Temp

[°C] Flow Rate

[Ton/h] Supply Temp

[°C] Return Temp

[°C]

FCW 54,609 40 47.14 54,609 40 47

SW 38,961 36 46.24 38,961 36 46

U 5485 W/m2-K 5485 W/m2-K

Heat Load 441 MW 448 MW It was finally concluded that no major modifications are required to handle the additional heat duty associated with removing seasonality in LNG production and heat duty from the GT inlet air cooling unit on the LP/MP-MR GT. Additional heat transfer area would be provided on the FCW/SW plate type exchangers by installing additional plates to the existing frame.

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RESTAGING OF THE PROPANE REFRIGERATION COMPRESSOR

A cross section of the propane compressor casing and rotor is illustrated in Figure 7. It is a straight through centrifugal unit with a horizontally split casing. Five impellers are mounted upon the 21.3 ft (6.5 m) long shaft. Three side loads add gas upstream of the suction at impellers 2, 3 and 4. The return channel and side load inlets are all vaned to align the downstream swirl angle. The impeller diffusers are all vaneless.

Figure 7. Train 3, 4 & 5 Propane Compressor

The previous section discussed the results of the liquefaction assessment study and concluded that, for Re-Rate Case 1, the bottleneck to maximising LNG output was the HHP (4th) stage of the propane compressor.

This stage had reached 1% from overload (stonewall) before the maximum available driver power had been consumed for the LP-MR/PR compressor string. It was therefore, also concluded that restaging of the compressor would be necessary to shift the 4th stage away from overload during normal operation.

Another consideration was a history of repeated impeller failures in this 4th stage. A root cause failure analysis was conducted in 2010 and reported in detail by (White et. al. 2011). It identified the root cause to be resonance induced high cycle fatigue of the impeller fundamental leading edge blade modal frequency. Operating within the vicinity of overload was concluded to be a contributing factor to exciting this mode. Three recommendations were made, two of which have since been implemented. The final recommendation, yet to be done, was to restage the compressor to shift the 4th stage away from overload.

Since centrifugal compressors are not normally intended to operate close to overload at normal suction and discharge conditions, there is clearly an historical reason why that is the case for this 4th stage. In fact, current operating conditions in all the compressor stages show that, on average, the mass flow rate has increased by 10% beyond the original normal design point listed in the API 617 datasheet released for procurement. Due to this fact, each stage operating point is constantly biased to the right hand side of the performance curve. The worst case is the 4th stage because it has the narrowest operating range. This is illustrated in Figure 8 (see Appendix) which is a plot of the current operating envelope, using bi-weekly data points taken between 2008 and 2009, against the design curve at rated speed. The historical reason behind the increased flow rates is project related and will not be elaborated on here. However, the associated risks were, at the time, largely unknown within the turbomachinery community. Eventually a paper written in 2006 by (Sorokes et. al.) reported that heavy gas compressors with large diameter impellers, having tall and slender blades at the leading edge, makes them particularly susceptible to mechanical integrity issues if operated in the vicinity of overload. They gave several examples of failed impellers, one of which bears a remarkable similarity to the present case. In addition, various other authors have since reported similar findings (Gresh 2006) & (Kammerer 2009).

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Physical Restrictions

The existing compressor design imposed certain restrictions (constraints) upon the degree of possible changes to the hardware to achieve the restage objectives. Before discussing them in detail, it should be understood that the restaging did not focus only on the problematic 4th stage. A reality for any multi-stage compressor with fixed suction and discharge conditions is that varying the duty of one stage will impact the duty of the other adjacent stages. The restaging study therefore, encompassed all four compressor stages.

Casing and Foundation The upper and lower casing design remained fixed, including the main suction and sideload nozzles. In addition, the casing support structure and foundation were fixed.

Drive End Shaft Diameter The drive end shaft is tapered to a final diameter of 10 inches. Any increase, to accommodate greater than design shaft power, would require a modified rigid coupling hub and also a larger bearing housing. It was concluded that the existing dimension should not be exceeded.

Shaft Length

By setting the compressor casing as a constraint, the rotor shaft length is also constrained by the position of the housed journal bearings. The first critical speed of the rotor Ncr_1 is influenced by the mass of the rotor and the overall rotor/bearing stiffness. Any redesign of the impellers, balance piston or thrust disc would thus require a rotordynamic study to validate the coupled shaft for correct operation.

Exit Volute Area

The exit volute area is limited axially by the discharge endwall position and radially by the balance piston. Avoidance of major redesign to the discharge endwall and nozzle necessitated keeping changes to the volute to a minimum.

Through Flow Area The suction volumetric flow rate to each compressor stage can be described by the following equation.

• 𝑄 = 𝜑𝑑3𝑁 𝜋2

4 (1)

Any change to the volumetric flow rate therefore, can generally be achieved by changing the flow coefficient, impeller diameter or shaft speed. In this case, the shaft speed is restricted and the flow coefficient may be assumed constant. Consequently, varying the impeller tip diameter is the main opportunity.

The maximum impeller tip diameter is limited by the blade stress at the exit for a given material. The Compressor OEM has specific impeller families from which to choose. They will cover a range of sizes (diameter and flow channel width) which limits the choice in the present case. In addition, increasing the impeller size usually requires an increase in shaft thickness to offset the larger rotating mass and inevitable reduction in critical shaft speed. The degree to which the existing shaft can be stepped up in diameter is of course, limited by the available techniques – either through the addition of material by welding or using metallic inserts.

The stage static flow channel (diffuser, return bend and return channel) must also be matched to the new volumetric flow rate whenever it has been changed. Generally, this is achieved by increasing the width of these components enough to ensure acceptable flow conditions are maintained. The constraints are primarily associated with the minimum required diaphragm material to safely withstand the applied axial force across each stage.

Performance Restrictions Required stage suction and discharge operating conditions were provided by the validated train simulation models (Re-Rate Cases 2 and 3) described in the previous section. Two design points were provided to the

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Compressor OEM; one for averaged summer and one for averaged winter conditions (Cases 2 and 3 respectively). Generally, the prescribed Winter Case differs from the Summer Case by a greater suction volumetric flow rate and lower polytropic head rise at each stage. This is due to a reduction in condenser pressure at the compressor discharge.

Standard design practice aerodynamic constraints were applied in the restage study, see for example (Drosjack et. al. 2011). This ensured that the stage loading was kept within safe limits. In the case of this compressor however, with its previous failure history, it was necessary to scrutinise the following four additional parameters linked to aerodynamic performance.

• Stage Flow Coefficient

• Stage Head Coefficient

• Impeller inlet relative Mach Number

• Impeller tip speed Plotting the latter three parameters against the first produces OEM experience (or ‘’cloud’’) diagrams for each stage. An indication of the risk involved with the rerated stage design can easily be obtained if all relevant previous successful designs are included.

Finally, the four compressor stage performance maps were subject to three main constraints commensurate with the standard API 617, 7th edition guidelines. Figure 9 (see Appendix) illustrates how these constraints are derived. They are applied to the rated operating point (winter point) and listed in Table 5 (below).

Table 5. Propane Compressor Stage Performance Map Constraints Applied at the Normal Design Point

Value % Tolerance %

Pressure rise to surge 4 -0Turndown to surge 12 -0Flow increase to overload 7 -0

Restage Results A cross sectional drawing of the compressor with marked changed components is given in Figure 10 (below). The green areas identify the changed rotating components. The blue areas identify the changed static components. It immediately strikes the eye that a significant amount of hardware, amounting to approximately 75% of the total static material volume within the outer casing, has been replaced to achieve the restage goals for the 4th stage only. This demonstrates the strong aerodynamic coupling between the individual compressor stages, which dictates that making design changes in one will affect the performance of the other stages to a varying degree. Other end users who are considering similar upgrades to their compressors should bear this fact in mind. Most often, a significant amount of hardware will require modification which invariably carries with it significant capital cost.

Impellers 3, 4 and 5 were replaced with impellers of the same family but with greater inlet and exit widths to accommodate the increased stage volumetric flow rate range. Impeller 4 was also trimmed by almost 15% at the exit diameter to achieve the required reduction in polytropic head. Impeller 5 actually required a 3% increase to its exit diameter to allow for its increased flow area of the same magnitude. For the same reason, the number of blades were also reduced from 15 to 13. Furthermore, the 5th impeller bore diameter was reduced by 7% which required machining of the existing shaft. Finally, the balance piston bore diameter was reduced to fit the lower shaft diameter behind the 5th impeller.

Stationary flow path components were replaced to enlarge the flow path area as required in the 3rd and 4th stages. The return channel wall of the 2nd stage was also replaced to reduce the number of static alignment vanes from 16 to 14. This was done to ensure adequate margin was maintained from modal excitation of the downstream impeller. The discharge volute was replaced even though its area remained the same. A larger flow path was created at the diffuser exit, volute tongue intersection by increasing the passage width.

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Figure 10. Cross Section of Restaged Compressor Table 6 (below) lists the percentage difference of suction volumetric flow rate, mass flow rate and polytropic head for each compressor stage. The Summer Case (Rerate Case 2) is compared to the original Normal Design operating point.

A re-distribution of stage loading is clearly evident as a result of the 9% increase in volumetric flow rate and 6% reduction in polytropic head of the 4th stage to increase its stonewall margin (flow increase to overload). Firstly, the corresponding reduction of volumetric flow rate in the 1st and 2nd stages essentially balances out the additional flow introduced to the 4th stage. Secondly, a significant increase in polytropic head was applied to the first three stages to partially offset the reduction in the 4th stage. Stage polytropic efficiency is not listed in Table 6 because it was within 0.5% of the original design value for each stage.

Overall, the required shaft power (including estimated parasitic losses) increased by 7%. This was primarily due to the large increase in mass flow rate introduced at the third sideload. It was confirmed that this additional power is within the existing compressor string maximum power rating and thus would require no further modifications to the hardware.

Table 6. Propane Compressor Rerate Cases 2 vs. Design

Stage 1 (impeller 1)

% Change of Summer Case (2) vs. Normal Design Point

Volumetric Flow Rate -6.0Mass Flow Rate -9.4Polytropic Head 4.6Stage 2 (impeller 2)Volumetric Flow Rate -3.4Mass Flow Rate -3.9Polytropic Head 3.2Stage 3 (impeller 3)Volumetric Flow Rate 0.5Mass Flow Rate 1.1Polytropic Head 8.1Stage 4 (impeller 4 & 5)Volumetric Flow Rate 9.2Mass Flow Rate 15.1Polytropic Head -5.5

Shaft Speed 0.2Overall Shaft Power 7.4

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Figure 11 (see Appendix) shows plots of the new stage performance maps for the restaged design. Both Summer and Winter Case maps are plotted, as full and dashed curves respectively. Their difference is a result of the different suction conditions at each stage which causes the map shift. The curves were normalised by the Summer Case point value of polytropic head and volumetric flow rate. The stage performance map margins, given as constraints in Table 5 are also shown in the legend of each plot. They are calculated for the Summer Case point and satisfy the constraints for all stages. It is noted however, that the increase to overload should also be considered for the Winter Case point where the volumetric flow rate is greatest. This value is highlighted in red in each plot. Again, the minimum constraint of 7%, from Table 5, has been satisfied at each stage. In particular, the problematic 4th stage now has a value of 13%, which is an increase of 12% beyond the existing design. This clearly demonstrates the success of the restaged design in achieving the set objectives without exceeding any of the prescribed physical or performance restrictions discussed previously. With the bottleneck at the 4th stage removed, there is now ample performance range within the propane compressor to permit full utilization of the available driver power during winter.

Figure 12 (see Appendix) shows the OEM’s experience diagrams that were previously mentioned, for the 3rd, 4th and 5th impellers. They confirm that all three aerodynamic parameters are within the current experience database.

The 4th impeller exit diameter cut-back of almost 15% was larger than previously experienced by the Compressor OEM. It was therefore, considered to pose significant risk to the intended performance of the compressor. The normal route to mitigating such risk in the detailed design phase is to perform a steady state CFD analysis. This was planned into the scope of work as an additional item and brought forward, as much as possible, in the project schedule to allow sufficient time to revise the design if required. Furthermore, the Company engaged a 3rd party expert consultancy to perform an independent design review.

The Compressor OEM was also tasked with performing lateral and torsional vibration analyses for the modified rotor. A full compressor string analysis was performed in accordance with API 617, 7th edition guidelines. Finally, a dynamic simulation of the restaged compressor was performed to assess the capability of the existing anti-surge control valves and line sizes. It was concluded that the existing system is satisfactory under the worst case scenarios that were analysed.

LP/MP-MR GAS TURBINE INLET AIR COOLING FOR POWER AUGMENTATION

The LP/MP-MR GT inlet air cooling duty sensitivity study for Train 4, identified that a chilling duty of 800 USRT was sufficient to meet the annual production target for Train 3, 4 and 5 of 4.8 MTPA.

Upon consideration it was evident that, GT inlet air cooling could also benefit LNG train reliability in the event of the helper motor tripping during summer operating conditions, where the average daily ambient temperature is high. Typically, in this scenario, the whole LNG train would also trip because the instantaneous compression load would not be reduced fast enough to prevent the GT from tripping on high exhaust temperature/maximum speed limit. This situation has actually occurred on various occasions over the past decade of operation with a negative impact on train availability. Prevention of the GT trip could be possible if the chilling duty were to be increased fast enough, by the required amount, in order to rebalance the compression load. A study is ongoing by several parties, including the GT OEM, to evaluate the feasibility of such a control philosophy. To the authors’ knowledge this would be the first of its kind for mechanical drive applications and would be a significant progression in maximizing reliability for all end users.

Inlet air cooling through heat exchange coils located within the intake filterhouse was selected as the most appropriate system over evaporative media cooling and fogging. The decision was founded upon the analysis of site specific historical meteorological data (hourly ambient dry bulb temperature and relative humidity) spanning the last decade, cooling water consumption, system reliability and a system economic analysis. Further information on the pros and cons of the different available cooling techniques can be found in the open literature. A particularly good description was given by (Al-Ibrahim et. al. 2010) for a site in Saudi Arabia.

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Figure 13 (below) illustrates the position of the chiller coils within the air inlet filterhouse. They are located on the clean air side of the conventional filtration media which consists of a coalescer followed by a pulse cleaning cartridge type filter. A drift eliminator is located immediately downstream of the chiller coils to prevent water condensate carry over into the GT axial compressor bellmouth. There are of course, various alternatives to how the existing filter house is retrofitted with chiller coils. Figure 13 (below) shows the simplest arrangement where the coil module is inserted behind the existing filtration stage by cutting and welding. The approximate site work was estimated to be 5 days, during which time, the GT would need to be shut down.

Figure 13. GT Air Inlet Filterhouse with Chiller Coils

The final retrofit design is subject to an ongoing study which is evaluating the most appropriate option. The main considerations are system reliability, maintainability, required shut down duration, system performance and overall lifetime cost.

The Frame 7EA output characteristics (power and heat rate) applied in the chiller system design were configured for the specific hardware in Train 5. This is because, the units in Train 5 were installed with Dry Low Nox (DLN1) combustion systems from commissioning in 2006. The units in Trains 3 and 4 are planned for retrofit with a DLN1 system by May 2013. In addition, the units have gone through various upgrades to the combustion and turbine hot gas path components, for example replacement of certain labyrinth seals with brush seals, which has slightly altered the GT performance characteristics.

Figure 14 (see Appendix) is a plot of GT output with a 7% power margin (assumed degradation) included. Four lines are plotted;

a. the baseline configuration

b. (a) plus the hardware upgrades mentioned above

c. (b) plus 10% power augmentation from inlet air cooling

d. (b) plus 25% power augmentation from inlet air cooling

The lines all span 7.2 deg C to 50 deg C, the lower limit being defined by the anti-icing protection at the GT axial compressor inlet. The latter two lines flatten off at approximately 118500HP (88.4MW) because the maximum swallowing capacity of the GT axial compressor has been reached, which defines the maximum power limit. In the second plot (Figure 15 – see Appendix), the GT axial compressor inlet temperature is plotted against the ambient temperature for lines (a), (c) and (d).

Drift Eliminator

Filtration Chiller Coils

Filter House

Extraction air

Silencer Panels

Ducting

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Design Options

At the start of the project, air cooled, refrigeration based or water cooled chilled water packages were all considered. The air cooled chilled water package required significant plot space to accommodate the fin-fan air coolers and the hot air from the fin-fans would possibly be an additional source of hot air recirculation to the GT inlets. This option was therefore, not selected. A refrigeration based chilling package (e.g. ammonia) was not selected due to complexity and hazardous nature of operating such a system. FCW was easily available and preliminary study indicated there is available spare capacity. Based on this consideration, it was decided to select a cooling water based chilled water package for GT inlet air cooling.

To provide inlet air cooling to the LP/MP-MR GT, a study was conducted to select between a centralised chilling unit or a local chilling unit using FCW. A centralised unit was sized to provide chilled water for three LNG trains from a central location and chilled water distributed to each GT unit by a chilled water pipe network. An option to install the chiller package locally to each GT unit was also assessed. This was carried out to optimise the chiller unit capacity against various constraints like availability of plot space, cost of pipe network, source of cooling water for the chiller package, etc.

Analysis and Results

Meteorological data of the Ras Laffan area recorded between 2000 and 2010, provided a design basis for determining the chilling load for chilling the ambient air down to a required GT inlet temperature. Hourly ambient (dry bulb) temperature and relative humidity data from January 1, 2000 to October 31, 2010 were applied in the analysis. Data for 2002 were omitted because of unavailability. A tabulation of the averaged data is given in Table 7 (see Appendix). The winter months (November – April) average ambient temperature is 21.1 deg C. Since power augmentation is required during summer months only, this number was taken as the temperature above which, chilling would be activated.

The enthalpy of air was calculated for all 82,068 data points using equation 2 below.

𝐻(𝑎) = 𝐶𝑝(𝑎)𝑇(𝑎) + (∆𝐻(𝑣) + 𝐶𝑝(𝑣)𝑇(𝑎))𝑋(𝑎) (2) The following procedure was adopted:

1. Determine saturated moisture content X(c) at chilled temperature T(c) using a psychometric equilibrium table.

2. Compare X(c) with X(a) (moisture content of ambient air).

3. If X(c) >= X(a), no water condensation occurs by chilling, and calculate enthalpy of chilled air using T(c) and X(a) using equation 2 above.

4. If X(c) < X(a), water condensation occurs by chilling, and calculate enthalpy of chilled air using T(c) and X(c) by using equation 2 above.

14

Table 8. Occurrence Percentiles for Enthalpy Data

In order to size the chiller correctly, the chilling duty was based upon the enthalpy difference between the ambient air T(a) and the GT inlet temperature after chilling T(c). This ensured that the amount of water condensate was correctly estimated for cases where the dew point was exceeded. The results for enthalpy difference and water condensation are shown in Figures 16 and 17 (see Appendix) for occurrence percentiles of 0.4%, 1% - 10% which are defined in Table 8 (above). The highest enthalpy difference line was not used for determining the chiller duty in order to avoid an extremely oversized chiller facility for 99.99% of the operation time. Within the industry, 0.4% occurrence percentile meteorological data are often used as a design basis of HVAC systems. However, for this project it was proposed to use the 10% difference data, because the smaller the size of chiller unit, the larger the feasibility of the chiller facilities in view of the available plot space, piping route space and utility supplies. Moreover, the 10 % occurrence percentile in this case would mean that the chiller capacity would not be enough for 10% of the time during the summer season. That is, about only 5% of annual hours but it would still satisfy about 60% of the chilling required in the 0.4% occurrence percentile.

The maximum condensation rate X(a) – X(c) was calculated by assuming that the ambient air at T(a) is saturated and then obtaining the corresponding X(a) from the enthalpy H(a), where H(a) = H(c) + ΔH and ΔH is the enthalpy difference used for the chiller duty. Table 9 (see Appendix) is a tabulation of the chiller duty results for the 0.4%, 1%, 2%, 3%, 4%, 5% and 10% occurrence percentiles. Three different centralised chiller design configurations are shown in the table, as discussed in the previous sub-section ; (A) - Air Chiller, (B) - Sea Water Cooled with 3 chiller packages and (C) - Sea Water Cooled with 4 chiller packages. The chiller duty, which increases from configuration A to C, is inversely proportional to the chilled temperature T(c) and thus GT power increment. As an example, configuration C achieves a GT inlet temperature reduction of 13.3 deg C for the 10% occurrence percentile in Table 9. The resulting chiller duty is 2,558 USRT and yields a GT power increment (determined from Figure 14) of 9.4%. The condensed water rate for this duty was calculated to be 7233 kg/h.

The FCW study concluded that there was spare capacity in the system for either centralised GT air inlet cooler package (one chiller package for three GT units) or a localised GT inlet air cooling package (one chiller package per GT unit). Typical plot plan arrangements for centralised and localised packages were developed to assess the piping routing complexity and the required modifications to the pipe racks for accommodating the new FCW/chilled water supply and return lines. A centralised package required significant modifications to existing pipe racks to accommodate larger pipe sizes but was more energy efficient. A localised package required a smaller foot print and minimum pipe runs for the FCW supply due to close proximity of the water cooled propane condensers. However, this option required more tie-ins and field

Enthalpy of Air ( kJ/kg dry air) per Ras Laffan Meteo Data (Jan. 2000 -- Oct. 201

Highest, Lowest and at Key Occurrence Percentiles

Occurrence Enthalpy (kJ/kg dry air)

Temperature(deg C)

Relative Humidity

(%)

Moisture Content(kg/kg dry air)

Highest Enthalpy 183.914 42.3 98.5 0.05485

0.4% OccurrenceEnthalpy 108.330 37.2 67.66 0.02759

1% Occurence Enthalpy 104.440 38.4 59.34 0.02562

2% Occurence Enthalpy 101.515 33.7 78.8 0.02639

3% Occurence Enthalpy 99.651 35.5 67.5 0.024926

4% Occurence Enthalpy 98.262 33.9 74.1 0.025041

5% Occurence Enthalpy 96.961 34.2 71.1 0.02441

10% Occurence Enthalpy 91.530 33.9 66.74 0.022425

15

work for three separate units. Taking into consideration that the proposed modifications would have to be carried out on an operating facility, a localised GT inlet air cooling package was selected to minimize construction risks and disruption to the operating facilities.

PROJECT SCHEDULE Implementation of any selected option for removing the seasonality impact is constrained by the LNG train scheduled shutdown and the shutdown period. The shutdown schedule is primarily driven by the GT maintenance schedule. The company currently follows a 16,000 operating hours schedule, where a sequence of Main Inspection, Combustion Inspection, Hot Gas Path Inspection and Combustion Inspection is employed in an eight year (64,000 operating hours) cycle. The first available opportunity will be in 2013 when Train 4 is scheduled for a Major Inspection with a shutdown window of approximately 28 days. At the start of the study, a high level schedule was drafted, Figure 18 (below), to help the project team to understand the schedule constraints and how much time would be allowed for analysing various options.

Figure 18. Project Schedule

CONCLUSIONS

A conceptual study to remove LNG production seasonality impact was completed in 14 months. The study involved five main parties (the Company, EPC contractor, Refrigeration Licensor, GT OEM and Compressor OEM) with the Company having overall management responsibility. This was accomplished through a kick off meeting and regular interaction using defined communication routes. The work became iterative in nature particularly between the Refrigeration Licensor, EPC Contractor and Compressor OEM. Clearly, it was in the Company’s best interests to minimise the number of iterations and therefore expended man-hours. A workflow plan with a roles and responsibility matrix was developed to establish the decision making process for all parties. Technical workshops were routinely held to ensure open communication between the various stakeholders and to track progress and clarifications.

It was concluded that summer production loss could be fully recovered with an investment of approximately $150M for three LNG trains. The required modifications to Trains 4 and 5 are the installation of localised GT inlet air cooling on the LP/MP/MR unit and restaging the propane compressor. For Train 3, only the localised GT inlet air cooling on the LP/MP-MR unit would be necessary.

The expected duration for procurement of the long lead items was estimated as 16 months. These consisted primarily of the propane compressor restaged rotor and stationary parts.

Modifications to the GT air inlet duct for installing inlet air cooling coils would not be on the shutdown critical path, as it requires only 5 days. The majority of the modifications required for the chilling package can be done offline without a shutdown. The critical path item would be the replacement of the propane rotor which requires approximately 20 days. The proposed execution schedule for the modification begins with Train 4 in 2013. Trains 3 and 5 have scheduled Hot Gas Path Inspections in 2015 and 2016 respectively. Since the normal HGPI shutdown duration is 11 days, an extension of approximately 9 days would be necessary on Train 5 to install the proposed modifications. Since this would have a major impact on train availability, an adjustment of the overall scheduled train shutdown plan would be necessary.

16

The project cost for three LNG trains, based on an unclassified cost estimate, is approximately $150M with an IRR of 33%. From a funding perspective, this was considered to be a feasible venture although further detailed analysis is omitted here. The project will serve to improve operational reliability and flexibility. This also provides capability of incremental LNG production to maintain LNG marketing targets which could have been impacted by unplanned outages.

NOMENCLATURE AGK = Al-Khaleej Gas Sales Train APCI = Air Products and Chemicals Incorporated API = American Petroleum Institute BCFD = Billion Standard Cubic Feet per Day CFD = Computational Fluid Dynamics DLN = Dry Low NOx EPC = Engineering Procurement Construction FCW = Fresh Cooling Water GT = Gas Turbine Engine HGPI = Hot Gas Path Inspection HP = High Pressure HHP = High High Pressure IAC = GT Inlet Air Cooling IRR = Internal rate of return LNG = Liquid Natural Gas LP = Low Pressure MCHE = Main Cryogenic Heat Exchanger MMTPA = Million Tonnes per Annum MR = Mixed Refrigerant MP = Medium Pressure MW = Mega Watt NOx = Nitrous Oxide OEM = Original Equipment Manufacturer PR = Propane SDC = Stream Day Capacity (production) SW = Sea Water TPH = Ton per Hour USRT = U.S Refrigeration Ton WHU = Waste heat recovery unit Symbols

Cp = Specific heat at constant pressure (kJ/kg deg C) d = Impeller exit diameter (m) H = Enthalpy (kJ/kg) N = shaft speed (rpm) Q = Volumetric Flow rate (m3/min) T = Temperature (deg C) X = Moisture content of air ΔH = Enthalpy difference (chiller duty) (kJ/kg) ΔH(v) = Latent heat of vaporization (kJ/kg) Φ = Stage flow coefficient Subscripts

a = Ambient dry air c = Chilled air v = Vapor

17

REFERENCES Al-Ibrahim, A.M., and Varnham A., “A review of inlet air-cooling technologies for enhancing the

performance of combustion turbines in Saudi Arabia”, Applied Thermal Engineering, Volume 30, Issues 14–15, October 2010, Pages 1879-1888.

Drosjack, M.J., Sorokes, J. M., and Miller, H.F. “Buying /Selling Serial Number 1” Proceedings of the 40th

Turbomachinery Symposium, Houston, 2011 Gresh, T. “Avoid Refrigeration Compressor Damage”. Article in Hydrocarbon Processing. Gulf Publishing

Company, Pages 93-94, October 2006 Issue. Kammerer, M. and Abhari, R. “Experimental Study on Impeller Blade Vibration During Resonance Part1:

Blade Vibration Due to Inlet Flow Distortion”. Journal. Engineering. Gas Turbines Power, Volume 131, Issue 2, March 2009.

Roberts M.J.,Bronfenbrenner J.C., Yu-Nan L. and Petrowski J.M. “Increased Capacity Single Train AP-XTM

Hybrid LNG Process”. GASTECH Conference, October 2002, Doha.

Sorokes, J. M., Miller, H. F. and Koch, J. M. “The Consequences of Compressor Operation in Overload”, Proceedings of the 35th Turbomachinery Symposium, Turbomachinery Laboratory, Texas A&M University, College Station, 2006

White, N. M., Laney, S., Zorzi, C. “RCFA for Recurring Impeller Failures in a 4.7 Mtpa LNG Train Propane

Compressor” Proceedings of the 40th Turbomachinery Symposium, Houston, 2011

ACKNOWLEDGEMENTS The authors would like to express their sincere thanks to the following people for their valuable contributions during this study:

Mr. Nafez Bseiso, Mr. Atul Dhokte, Mr. Meshal Al-Mohannadi, Dr. Basel Wakileh, Mr. Jamal Al-Anbari, Mr. Mohammed Shikari, Mr. Fahad Al-Khater, Mr. Abhay Nafde, Mr. Haridas Balki and Mr. Atul Deshpande of RasGas; Mrs. Alicia Pollard, Mr. David Bucci and Mr. Robert Snyder of Elliott Group; Mr. Robert Saunderson and Mr. Joseph Wehrman of APCI; Mr. Robert Johnston of GE and Mr. Kinya Iwabuchi, Mr. Yuji Sato, Mr. Kuniteru Tomioka and Mr. Masataka Ryushi of Chiyoda.

18

APPENDIX Figures :

Main Cryogenic Heat Exchanger

Helper Motor (12 MW)

FCWFeed Gas

FCW

MR

Helper Motor (12 MW)

LNG

Frame 7 GT/MR(75~85 MW)

MR/1 MR/2

Frame 7 GT/PR(75~85 MW)

PR MR/3

FCW

181 MW, 19,351 tph FCW

74 MW, 7412 tph FCW

Figure 2. Train 3, 4 and 5 Arrangement

Figure 3. Facility Plot Plan Showing Sources of Hot Air

Helper Motor (0 - 12 MW)

Helper Motor (0 - 12 MW)

PR HP-MR Frame 7EA (84 - 67 MW)

Frame 7EA (84 - 67 MW)

LP-MR MP-MR

19

Figure 8. Train 3, 4th Stage Head and Work vs. Volumetric Flow Rate Bi-Weekly Operating Points

Taken Between 2008 - 2009

Figure 9. Stage Performance Map Margin Definitions

Figure 11. Stage Performance Maps for Restaged

0.75

0.80

0.85

0.90

0.95

1.00

1.05

1.10

0.80 0.85 0.90 0.95 1.00 1.05 1.10 1.15 1.20

Nor

mal

ised

Pol

y H

ead

[-]

Normalised Volumetric Flow Rate [-]

Stage 1 110% Speed

100% Speed

99% Speed

98% Speed

97% Speed

96% Speed

95% Speed

Winter Point

Summer Point

Full lines (Summer Case)Dashed lines(Winter Case)

Press Rise to Surge = 4.4%

Turndown to Surge = 12.9%

Flow Increase to Overload = 16.3%,10.5%

0.70

0.75

0.80

0.85

0.90

0.95

1.00

1.05

1.10

0.80 0.85 0.90 0.95 1.00 1.05 1.10 1.15 1.20

Nor

mal

ised

Pol

y H

ead

[-]

Normalised Volumetric Flow Rate [-]

Stage 2 101% Speed

100% Speed

99% Speed

98% Speed

97% Speed

96% Speed

95% Speed

Winter Point

Summer Point

Full lines (Summer Case)Dashed lines (Winter Case)

Press Rise to Surge = 5.7%

Turndown to Surge = 12.1%

Flow Increase to Overload = 15.5%, 7.3%

0.60

0.65

0.70

0.75

0.80

0.85

0.90

0.95

1.00

1.05

1.10

0.70 0.80 0.90 1.00 1.10 1.20 1.30

Nor

mal

ised

Pol

y H

ead

[-]

Normalised Volumetric Flow Rate [-]

Stage 3 101% Speed

100% Speed

99% Speed

98% Speed

97% Speed

96% Speed

95% Speed

Winter Point

Summer Point

Full lines (Summer Case)Dashed lines (Winter Case)

Press Rise to Surge = 3.7%

Turndown to Surge = 14.0%

Flow Increase to Overload = 22.6%, 13.1%

0.60

0.65

0.70

0.75

0.80

0.85

0.90

0.95

1.00

1.05

1.10

0.70 0.80 0.90 1.00 1.10 1.20 1.30

Nor

mal

ised

Pol

y H

ead

[-]

Normalised Volumetric Flow Rate [-]

Stage 4 101% Speed

100% Speed

99% Speed

98% Speed

97% Speed

96% Speed

95% Speed

Winter Point

Summer Point

Full lines (Summer Case)Dashed lines (Winter Case)

Press Rise to Surge = 4.0%

Turndown to Surge = 12.5%

Flow Increase to Overload = 21.3%, 13.4%

Potentially in stonewall

20

Figure 12. OEM Experience Plots for 4th Stage (Restaged Design)

Figure 14. GT Power Output vs. Ambient Temperature Figure 15. GT Inlet Temperature vs. Ambient Temperature

21

Figure 16. Enthalpy Difference for Occurrence Percentiles

Figure 17. Water Condensation for Occurrence Percentiles

22

Tables :

Table 2. Train 4 Re-Rating Cases 1, 2 and 3 with Modified Propane Compressor and Inlet Air Cooling On LP/MP-MR Turbine

Case MR Power Required

MW

MR Power Available

MW

Power from IAC

MW

IAC , USRT

PR power

Required MW

PR Power Available MW Comments

Summer 84.30 80.26 4.04 800 78.14 79.55 3636 rpm

Minimum LNG required to remove seasonality

Summer -max

(August) 84.67 80.26 4.40 960 78.34 79.55

3636 rpm Maximum LNG with 960

USRT Chiller

Winter 88.16 88.94 0 None 64.8 85.50 3534 rpm Current Winter LNG rate

Winter- max (May) 88.26 82.49 5.77 960 74.26 81.18

3616 rpm

LNG production limited by 7% stonewall margin on Propane

Compressor

Extreme Summer 76.17 73.81 2.36 800 69.58 70.31

3551 rpm

LNG production limited by power available to Propane-

HPMR compressor.

Table 3. Train 3 Re-Rating Case 4 with Current Propane Compressor and Inlet Air Cooling On LP/MP-MR Turbine

Case

MR Required,

MW

MR Power Available

MW

Power from IAC

MW

IAC , USRT

PR power

Required MW

PR Power Available MW Comments

Summer 85.51 81.31 4.20 800 78.07 81.11 3581 rpm

LNG production limited by 7% stonewall margin on

Propane Compressor and power available from 800

USRT chiller

Winter 88.41 83.58 4.83 800 70.75 81.39 3518 rpm

Summer 87.50 82.70 4.80 800 74.62 81.53 3550 rpm

Summer 87.43 82.85 4.58 800 77.86 82.35 3570 rpm

23

Table 7. Averaged Meteorological Data

Ras Laffan Area Ambient Temperature Averages (Unit: deg C) Attachment-1 Rev. 1

Highest Temp

Average Temp

Lowest Temp

Highest Temp 23.4 24.6 NR 24.8 31.1 25.9 21.3 27.4 22.5 27.0 25.3Average Temp 18.1 16.1 NR 19.4 19.3 17.7 15.4 15.6 15.4 18.5 17.3Lowest Temp 10.0 8.6 NR 11.8 8.4 9.0 7.8 6.7 8.4 9.3 8.9Highest Temp 22.7 27.5 NR NR 24.3 27.3 29.9 29.1 32.9 33.5 28.4Average Temp 17.2 17.2 NR NR 17.1 19.2 18.1 16.6 18.9 20.0 18.0Lowest Temp 10.9 8.0 NR NR 8.4 11.0 10.6 9.0 10.5 9.7 9.8Highest Temp 32.3 24.1 35.1 32.1 31.4 33.1 30.6 33.7 30.6 34.1 31.7Average Temp 21.0 20.0 20.7 21.3 21.0 21.4 20.8 22.2 20.6 23.1 21.2Lowest Temp 12.3 14.5 10.3 12.2 12.9 12.1 12.0 10.1 11.7 13.3 12.1Highest Temp 40.3 37.0 38.2 34.0 36.2 37.4 38.5 39.3 42.8 35.7 37.9Average Temp 27.1 25.9 25.6 24.9 25.1 25.8 26.8 25.9 34.5 26.8 26.8Lowest Temp 12.5 15.2 17.7 16.4 15.4 14.2 16.3 17.1 26.7 17.3 16.9Highest Temp 41.5 42.3 41.5 41.1 40.3 41.4 43.3 42.6 40.9 42.9 41.8Average Temp 30.7 30.8 30.6 29.6 29.3 31.4 31.9 31.6 31.6 31.0 30.9Lowest Temp 21.1 19.8 19.6 19.9 16.8 23.7 21.4 23.7 23.2 23.0 21.2Highest Temp 42.0 42.4 43.9 42.6 41.9 43.3 43.7 45.4 43.5 46.3 43.5Average Temp 31.7 31.9 33.5 32.1 32.4 33.1 33.2 33.4 32.8 33.3 32.7Lowest Temp 21.9 20.6 24.0 21.4 22.7 23.9 24.6 23.9 21.1 26.0 23.0Highest Temp 47.1 45.5 41.5 43.5 42.3 41.9 44.4 45.1 42.8 46.4 44.1Average Temp 34.7 33.7 33.2 33.8 33.3 34.4 34.4 34.5 34.5 34.8 34.1Lowest Temp 26.8 22.9 25.2 25.3 26.0 27.2 27.4 26.5 26.5 26.8 26.1Highest Temp 45.6 44.3 43.5 43.5 43.8 44.1 43.0 42.3 45.0 49.9 44.5Average Temp 34.8 34.1 33.8 33.1 33.8 34.1 34.5 34.3 34.3 35.3 34.2Lowest Temp 27.6 26.1 26.3 18.6 26.1 27.9 28.1 27.2 27.1 28.8 26.4Highest Temp 41.3 44.1 43.5 41.4 41.2 40.1 42.4 43.4 41.5 42.1 42.1Average Temp 31.2 32.5 33.8 31.3 31.6 32.0 32.9 32.8 32.3 33.5 32.4Lowest Temp 23.3 22.8 26.3 22.0 23.4 22.4 25.8 24.4 24.0 25.7 24.0Highest Temp 37.7 41.5 40.7 39.5 36.6 37.7 37.0 37.7 37.2 39.4 38.5Average Temp 28.6 28.8 29.2 27.2 28.6 29.5 29.8 29.1 28.9 30.2 29.0Lowest Temp 20.6 20.2 20.6 18.4 20.4 22.0 22.1 20.6 21.6 21.1 20.8Highest Temp 33.3 34.7 32.2 32.5 34.5 35.9 33.8 28.9 33.9 NR 33.3Average Temp 22.7 23.3 23.6 24.8 24.7 24.2 24.5 23.1 25.4 NR 24.0Lowest Temp 15.8 14.0 15.2 16.1 17.3 15.6 14.2 16.4 16.1 NR 15.6Highest Temp 30.1 29.7 27.2 25 28.7 23.0 29.0 25.6 26.1 NR 27.2Average Temp 18.8 22.1 19.6 18.1 20.5 16.1 19.2 17.6 19.7 NR 19.1Lowest Temp 9.5 12.6 12.5 10.1 12.2 8.8 9.5 9.1 13.8 NR 10.9

2007 2008 2009 201010 Year AverageKind of

Temperature 2006Month 2000 2001 2003 2004 2005

July

August

September

October

November

December

January

February

March

April

May

June

24

Table 9. Analysis Results for Chiller Duties

Case No.

Case Description

Design Ambient Conditions

1D

esign Conditions

1.1 D

ry Bulb tem

perature42.3

degC37.2

degC38.4

degC33.7

degC35.5

degC33.9

degC34.2

degC33.9

degC1.2

Relative H

umidity

98.5 %

67.7 %

59.3 %

78.8 %

67.5 %

74.1 %

71.1 %

66.7 %

1.3 M

oisture0.05485

kg/kg dry0.02759

kg/kg dry0.02562

kg/kg dry0.02639

kg/kg dry0.02493

kg/kg dry0.02504

kg/kg dry0.02441

kg/kg dry0.02243

kg/kg dry1.4

Dew

Point

42.2 degC

30.3 degC

29.0 degC

29.5 degC

28.5 degC

28.4 degC

28.4 degC

26.8 degC

1.5 E

nthalpy183.91

kJ/kg dry108.33

kJ/kg dry104.44

kJ/kg dry101.52

kJ/kg dry99.65

kJ/kg dry98.26

kJ/kg dry96.96

kJ/kg dry91.53

kJ/kg dry1.6

GT O

utput with 7%

power m

argin89,500

HP92,800

HP92,000

HP95,200

HP94,500

HP94,800

HP94,600

HP94,800

HP

AAir Cooled Chiller Plant

A2

Com

pressor Inlet Chilled A

ir Condition

A2.1

Dry B

ulb temperature

28.6 degC

28.6 degC

28.6 degC

28.6 degC

28.6 degC

28.6 degC

28.6 degC

28.6 degC

A2.2

Relative H

umidity

100.0 %

100.0 %

100.0 %

100.0 %

99.6 %

98.1 %

98.7 %

92.3 %

A2.3

Moisture

0.02496 kg/kg dry

0.02496 kg/kg dry

0.02496 kg/kg dry

0.02496 kg/kg dry

0.02487 kg/kg dry

0.02449 kg/kg dry

0.02464 kg/kg dry

0.02303 kg/kg dry

A2.4

Dew

Point

28.6 degC

28.6 degC

28.6 degC

28.6 degC

28.54 degC

28.31 degC

28.40 degC

27.31 degC

A2.5

Enthalpy

92.49 kJ/kg dry

92.49 kJ/kg dry

92.49 kJ/kg dry

92.49 kJ/kg dry

92.26 kJ/kg dry

91.29 kJ/kg dry

91.67 kJ/kg dry

87.56 kJ/kg dry

A2.6

Mass Flow

277.21 kg/s

277.21 kg/s

277.21 kg/s

277.21 kg/s

277.19 kg/s

277.08 kg/s

277.12 kg/s

276.69 kg/s

A2.7

Mass Flow

dry270.46

kg/s270.46

kg/s270.46

kg/s270.46

kg/s270.46

kg/s270.46

kg/s270.46

kg/s270.46

kg/s

A3

Chiller Load @

Max A

mbient C

onditionsA

3.1 E

nthalpy difference91.43

kJ/kg dry15.84

kJ/kg dry11.95

kJ/kg dry9.03

kJ/kg dry7.40

kJ/kg dry6.98

kJ/kg dry5.29

kJ/kg dry3.97

kJ/kg dryA

3.2 C

hilling Duty

89,020,326kJ/h

15,427,511kJ/h

11,639,989kJ/h

8,792,045kJ/h

7,200,835kJ/h

6,792,870kJ/h

5,153,337kJ/h

3,866,867kJ/h

Chilling D

uty24,727.9

kW/h

4,285.4kW

/h3,233.3

kW/h

2,442.2kW

/h2,000.2

kW/h

1,886.9kW

/h1,431.5

kW/h

1,074.1kW

/h Chilling Duty

7,032 USRT

1,219 USRT

920 USRT

695 USRT

569 USRT

537 USRT

407 USRT

305 USRT

A4

GT O

utputA

4.1 G

T Output w

ith 7% pow

er margin

98,500 HP

98,500 HP

98,500 HP

98,500 HP

98,500 HP

98,500 HP

98,500 HP

98,500 HP

A4.2

GT O

utput Increment

9,000 HP

5,700 HP

6,500 HP

3,300 HP

4,000 HP

3,700 HP

3,900 HP

3,700 HP

A5

Condensed W

ater Rate

29,103kg/h

2,561kg/h

643kg/h

1,392kg/h

55kg/h

536kg/h

0kg/h

0

BSea W

ater Cooled Chiller Plant (3 chiller packages)B

2C

ompressor Inlet C

hilled Air C

onditionB

2.1 D

ry Bulb tem

perature26.1

degC26.1

degC26.1

degC26.1

degC26.1

degC26.1

degC26.1

degC26.1

degCB

2.2 R

elative Hum

idity100.0

%100.0

%100.0

%100.0

%100.0

%100.0

%100.0

%100.0

%B

2.3 M

oisture0.02150

kg/kg dry0.02150

kg/kg dry0.02150

kg/kg dry0.02150

kg/kg dry0.02150

kg/kg dry0.02150

kg/kg dry0.02150

kg/kg dry0.02150

kg/kg dryB

2.4 D

ew P

oint26.1

degC26.1

degC26.1

degC26.1

degC26.1

degC26.1

degC26.1

degC26.1

degCB

2.5 E

nthalpy81.04

kJ/kg dry81.04

kJ/kg dry81.04

kJ/kg dry81.04

kJ/kg dry81.04

kJ/kg dry81.04

kJ/kg dry81.04

kJ/kg dry81.04

kJ/kg dryB

2.6 M

ass Flow279.40

kg/s279.40

kg/s279.40

kg/s279.40

kg/s279.40

kg/s279.40

kg/s279.40

kg/s279.40

kg/sB

2.7 M

ass Flow dry

273.52 kg/s

273.52 kg/s

273.52 kg/s

273.52 kg/s

273.52 kg/s

273.52 kg/s

273.52 kg/s

273.52 kg/s

B3

Chiller Load @

Max A

mbient C

onditionsB

3.1 E

nthalpy difference102.87

kJ/kg dry27.29

kJ/kg dry23.40

kJ/kg dry20.47

kJ/kg dry18.61

kJ/kg dry17.22

kJ/kg dry15.92

kJ/kg dry10.49

kJ/kg dryB

3.2 C

hilling Duty

101,296,187kJ/h

26,870,738kJ/h

23,040,364kJ/h

20,160,198kJ/h

18,324,770kJ/h

16,957,060kJ/h

15,676,002kJ/h

10,328,248kJ/h

Chilling D

uty28,137.8

kW/h

7,464.1kW

/h6,400.1

kW/h

5,600.1kW

/h5,090.2

kW/h

4,710.3kW

/h4,354.4

kW/h

2,869.0kW

/h Chilling Duty

8,002 USRT

2,123 USRT

1,820 USRT

1,593 USRT

1,448 USRT

1,340 USRT

1,238 USRT

816 USRT

B4

GT O

utputB

4.1 G

T Output w

ith 7% pow

er margin

100,000HP

100,000HP

100,000HP

100,000HP

100,000HP

100,000HP

100,000HP

100,000HP

B4.2

GT O

utput Increment

10,500HP

7,200HP

8,000HP

4,800HP

5,500HP

5,200HP

5,400HP

5,200HP

B5

Condensed W

ater Rate

32,839kg/h

5,997kg/h

4,057kg/h

4,815kg/h

3,373kg/h

3,487kg/h

2,865kg/h

911kg/h

CSea W

ater Cooled Chiller Plant (4 chiller packages)C

2C

ompressor Inlet C

hilled Air C

onditionC

2.1 D

ry Bulb tem

perature20.6

degC20.6

degC20.6

degC20.6

degC20.6

degC20.6

degC20.6

degC20.6

degCC

2.2 R

elative Hum

idity100.0

%100.0

%100.0

%100.0

%100.0

%100.0

%100.0

%100.0

%C

2.3 M

oisture0.01526

kg/kg dry0.01526

kg/kg dry0.01526

kg/kg dry0.01526

kg/kg dry0.01526

kg/kg dry0.01526

kg/kg dry0.01526

kg/kg dry0.01526

kg/kg dryC

2.4 D

ew P

oint20.6

degC20.6

degC20.6

degC20.6

degC20.6

degC20.6

degC20.6

degC20.6

degCC

2.5 E

nthalpy59.46

kJ/kg dry59.46

kJ/kg dry59.46

kJ/kg dry59.46

kJ/kg dry59.46

kJ/kg dry59.46

kJ/kg dry59.46

kJ/kg dry59.46

kJ/kg dryC

2.6 M

ass Flow284.68

kg/s284.68

kg/s284.68

kg/s284.68

kg/s284.68

kg/s284.68

kg/s284.68

kg/s284.68

kg/sC

2.7 M

ass Flow dry

280.40 kg/s

280.40 kg/s

280.40 kg/s

280.40 kg/s

280.40 kg/s

280.40 kg/s

280.40 kg/s

280.40 kg/s

C3

Chiller Load @

Max A

mbient C

onditionsC

3.1 E

nthalpy difference124.46

kJ/kg dry48.87

kJ/kg dry44.98

kJ/kg dry42.06

kJ/kg dry40.19

kJ/kg dry38.81

kJ/kg dry37.50

kJ/kg dry32.07

kJ/kg dryC

3.2 C

hilling Duty

125,632,608kJ/h

49,335,095kJ/h

45,408,374kJ/h

42,455,762kJ/h

40,574,166kJ/h

39,172,053kJ/h

37,858,772kJ/h

32,376,503kJ/h

Chilling D

uty34,897.9

kW/h

13,704.2kW

/h12,613.4

kW/h

11,793.3kW

/h11,270.6

kW/h

10,881.1kW

/h10,516.3

kW/h

8,993.5kW

/h Chilling Duty

9,925 USRT

3,897 USRT

3,587 USRT

3,354 USRT

3,205 USRT

3,094 USRT

2,991 USRT

2,558 USRT

C4

GT O

utputC

4.1 G

T Output w

ith 7% pow

er margin

103,700HP

103,700HP

103,700HP

103,700HP

103,700HP

103,700HP

103,700HP

103,700HP

C4.2

GT O

utput Increment

14,200HP

10,900HP

11,700HP

8,500HP

9,200HP

8,900HP

9,100HP

8,900HP

C5

Condensed W

ater Rate

39,964kg/h

12,446kg/h

10,458kg/h

11,235kg/h

9,757kg/h

9,873kg/h

9,236kg/h

7,233kg/h

Enth

alpy Base

Lin

e: dry air 0

degC

1 U

SR

T =

3.5

163kW

= 3

,024kC

al/h =

0.6

05 m

3/h C

hille

d Wate

r x (12-7 =

5) de

gC te

mpe

rature

deffe

rence, w

hic

h is a stan

dard of w

ater c

hillin

g indu

stry.

GE 7EA G

T Inlet Air Chilling: C

hiller Load(a

t Hig

h E

nth

alp

y D

iffere

nce C

ases)

Parameters

Highest Enthalpy

Attachment-6

Rev. 2

10% O

ccurrenceEnthalpy

1% O

ccurrenceEnthalpy

2% O

ccurrenceEnthalpy

5% O

ccurrenceEnthalpy

0.4% O

ccurrenceEnthalpy

3% O

ccurrenceEnthalpy

4% O

ccurrenceEnthalpy