Allen 5000 series

A completely new medium speed engine was launched in 1998, the 320 mm bore/410 mm stroke Allen 5000 series having an output of 525 kW/cylinder to cover a power range from 3150 kW to 10 500 kW at 720/750 rev/min from in-line six and eight and V10, 12, 14, 16, 18 and 20-cylinder models. The V12 and V16 versions were released in 1998 (Figure 17.5), the other configurations planned to follow

Cylinder Engine Cross Section
Figure 17.5 Cross-section of Allen 5000 series engine

progressively. The initial applications focused on land power generation but marine propulsion and auxiliary power opportunities were anticipated later.

With no carry over of components from existing Allen engines, the brief was to produce a robust design that was economical to manufacture, incorporated multi-functional components, had the minimum number of external pipes, used hydraulic tensioning for critical fasteners, and was easy to service in the field. Excellent fuel economy, smoke-free combustion and low emissions of other pollutants were also targeted during the design process.

The crankshaft design is typical for a medium speed engine: a one-piece steel slab forging, hardened and tempered; the material has an ultimate tensile strength of 900 Mpa; and the front end and flywheel end flanges are forged integrally with the shaft. In order to minimize the cylinder centre distance a crankpin diameter of 310 mm and a main journal diameter of 330 mm were specified, together with recessed fillet radii for both crankpins and journals. A torsional vibration analysis of the six- to-V16-cylinder versions was carried out, indicating that a crankshaft torsional vibration damper would be required, the choice of a tuned or viscous damper depending on the cylinder configuration. As a result of this analysis a cylinder bank angle of 52 degrees was chosen for the subsequent calculations.

Both classical and finite element stress analyses for the crankshaft were executed. Safety factors of 1.7 or greater under combined bending and torsional loads were predicted for engines used in fixed speed generating applications. The external imbalance of the in-line six, V12 and V16 models is zero for all primary and secondary forces and couples. The in-line nine, V10 and V14 models all have non-zero primary couples. The V10-14 models show inherent vertical and horizontal couples increasing as the bank angle is narrowed. By applying a suitable counter-weighting arrangement these couples can be partially balanced to leave a minimum possible imbalance. Further reduction is possible through the use of balancer shafts or reductions in reciprocating mass. In order to achieve the optimum balance for the V10 and V14 engines, extra balance weight must be applied to the crankshaft to counter both the primary rotating and mean primary reciprocating couples. As a result of the detailed analysis, the 52 degree bank angle was confirmed as the best choice for the V-cylinder engines.

Bi-metallic aluminium tin bearings were selected for both main and large end bearings. The crankpin diameter of 310 mm made it impossible to design a two-piece connecting rod arrangement that could be withdrawn up the cylinder bore (320 mm). A three-piece rod design of the marine-type was therefore selected; this enables the piston and rod to be removed through the liner while leaving the big end assembly bolted undisturbed on the crankpin. An advantage of this rod type is that the overhaul height requirement is less than for a two-piece design.

All the joint faces of the connecting rod are machined at 90 degrees to the longitudinal axis of the rod. No serrations are used and the three parts of the assembly are located to each other with fitted dowels arranged to prevent incorrect assembly. The shank of the rod is bolted to the big end block with four 24 mm studs, and the two halves of the big end block are secured with two 39 mm studs; both sets of fasteners are hydraulically tensioned. The majority of the surfaces on the rod remain in the 'as forged' condition. A full finite element analysis of the connecting rod showed that, apart from an area around the small end of the rod, the safety factors were generous. More material was added to the small end, this change increasing the safety factor at this point to an acceptable value.

The cylinder head is a rigid design of sufficient depth to evenly distribute the clamping load from the four hydraulically-tensioned head studs over the gas sealing face between the head and the cylinder liner; it is cast in compacted graphite cast iron. The MAGMASOFT casting simulation program was used to optimize the casting process. The inlet and exhaust ports are located on the same side of the head, with the inlet ports designed to impart a low level of swirl.

Both inlet and exhaust valve seats are intensively cooled, as is the rest of the flame face of the head. The cooling strategy of both head and liner was optimized using the VECTIS CFD simulation program, and a complete model of the coolant passages was made. The flow regime was evaluated throughout the model, with particular attention given to ensure there was a positive water flow velocity of between 1.5 m/s and 3 m/s around all the valve seats and drillings in the flame face of the head. As a result of this optimized cooling, the predicted temperatures at the full load condition are evenly distributed around the combustion face. These temperature values, together with stresses and safety factors, were calculated from a full 3D finite element model of the head. After some minor refinements to the design to increase the stiffness of the head stud bosses, all the temperatures, stresses and safety factors were within the capabilities of the compact graphite material.

At the conceptual phase of the design it was decided that the cylinder liner should be strategically cooled, and several different layouts were analysed using finite element techniques to calculate the temperature profile in the upper region of the liner. Layouts 1 and 2 (Figure 17.6) were rejected on the grounds that the predicted temperature at the

top ring reversal was too high or that the temperature profile was unsatisfactory. Layout 3, which uses a conventional bore cooling arrangement, was selected on the basis that the predicted liner surface temperature at the top ring reversal point was 169°C while 100 mm down from that point it was 138°C.

The liner is made from centrifugally cast iron. An anti-polishing ring is fitted at its top to eliminate bore polishing and ensure both a low and stable lube oil consumption over extended engine operation. The coolant flow around the liner was optimized using the VECTIS CFD simulation code. The coolant passages of both liner and cylinder head were incorporated into a single model.

A two-piece piston design, with steel crown and nodular cast iron skirt, was considered the best choice for securing a robust long life when operating at high cylinder pressures (210 bar design limit). The crown is cooled by oil supplied via drillings in the connecting rod and piston pin. Three compression rings and one oil control ring are fitted and, in conjunction with the anti-polishing ring in the liner, are expected to ensure low and stable lube oil consumption over a long period, especially in heavy fuel operation.

The crankcase is a one-piece nodular iron casting comprising the cylinder blocks and crankcase. The crankshaft is underslung and the main bearing caps are retained by hydraulically-tensioned vertical studs and horizontal side bolts. A gear case is incorporated at both ends of the crankcase, that at the flywheel end having an extra main bearing to support the mass of the flywheel. The air manifold is formed in the space in the centre of the vee. At the bottom of the crankcase are substantial mounting rails which impart additional stiffness in that area. Many of the oil and water transfer passages are either cast in or drilled to minimize the number of pipes used.

A full 3D finite element analysis of the V12-cylinder crankcase was carried out to confirm the design goals of high rigidity in both bending and torsion, together with modest stress levels, to ensure trouble-free operation during the life of the engine. Before casting the first examples of the crankcase, the MAGMASOFT casting simulation program was used to optimize the design of the runners and risers, and to confirm that the crankcase design was such that good quality castings would be produced without the need for corrections.

A conventional push rod operated valve gear was selected. The camshaft is mounted high in the cylinder block to minimize the length of the push rods and keep the valve gear stiffness high, and also to keep the length of the high pressure fuel injection pipe as short as possible. The camshaft is created from individual cylinder sections, which are then bolted to cylindrical sections (Figure 17.7). Such a design allows single sections of the camshaft to be removed sideways from the engine; it also enables solid bearing housings to be used, improving the stiffness of the cylinder block in that area. The proportions of the camshaft segments are deliberately generous to ensure moderate fuel cam operating stresses and minimum bending and torsional deflections when operating at the full design pressure of the fuel injection system (2000 bar). The camshaft is driven by a train of spur gears mounted at the flywheel end of the engine.

In specifying the fuel injection system, it was believed that fuel tolerance and injection flexibility to manage the conflicting requirements of low fuel consumption and low emissions could only be achieved by using digital electronics. R&D by Bosch (Hallein) removed the need for mechanical systems to alter the entire camshaft or parts that act on the fuel cams by exploiting proven digital electronic technology in a configuration co-developed by Rolls-Royce and Bosch. High efficiency and low emissions at all loads are promised, each cylinder calibrated for optimum performance. Individual pump and injector units for each cylinder are served by a side entry double-skinned high pressure fuel pipe. The pumping system is based on a conventional arrangement of camshaft and roller-operated plunger providing injection pressure capability over 2000 bar, with injection controlled electronically by a solenoid servo system via a spill valve.

By applying Ricardo engine cycle simulation software (WAVE) during both conceptual and definitive design phases, the geometries and thermodynamics of the engine performance were continually developed to allow full optimization of the power cycle objectives. ABB Turbo Systems' high efficiency TPL axial turbochargers are used. The charge air system comprises an integrated casing onto which the turbocharger can be mounted, and allows a high performance intercooler to be fitted within the casting. Two-stage intercooling is applied to provide charge air heating at lower loads and give improved heat recovery when this is required.

See Medium Speed Engines Introduction chapter for the evolution of the Allen 5000 series design parameters.

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