High-Speed High-Output Diesel Engines

High-Speed High-Output Diesel Engines - 35 Years of Development of Railroad and Marine Applications

By Markus von Kienlin and G. W. Maybach

History

TO THE EXPERT, the name of Maybach Motorenbau immediately recalls a special class of engines characterized by the following outstanding features: high speed, light weight, small size, high fatigue strength (due to special design and production measures,) and low operating cost.

The unique development of the Zeppelin airship after the historic accident at Echterdingen in 1908, was decisively governed by the fact that Wilhelm Maybach, Gottlieb Daimler's collaborator, made available to Count Zeppelin an engine designed by his son, Karl Maybach. This was a 6-cyl 150 hp 1200 rpm engine with a fuel consumption of 254 g/hph and a weight-to-power ratio of no more than 2,99 kg/hp, in contrast to the engine installed in the first Zeppelin airship which had a weight of 26 kg/hp and a consumption of 508 g/hph. For the production of the new engine, Count Zeppelin and Wilhelm Maybach founded, on March 23, 1909, the Luftfahrzeug-Motorenbau G.m.b.H. which was affilleated to Luftschiffbau Zeppelin. The Maybach Motorenbau, which was entrusted to Karl Maybach from the outset, produced the engines for further Zeppelin airships and for aeroplanes. Among those engines, as early as 1917, highly rated high-compression high-altitude engines were built. Fig. 1 shows such a high-altitude aircraft engine. Fig. 2 is a sectional view of the cylinder liner and wrist pin area.

After the enforced interruption of almost the entire production at the end of World War I, new and related lines of production suitable for the skilled staff had to be found. The choice included Otto and diesel engines for a number of purposes. Research continued on Zeppelin airship engines including the well-known 550 hp VL-type engine (Fig. 3) for the airship "Graf Zeppelin," which accomplished the famous world tours under the command of the late Dr. Eckener, who until his death in 1954 was chairman of the Maybach Motorenbau. The first powerful high-speed diesel engine was developed in the Maybach workshops between 1919 and 1923. Construction of this G 4 engine (Fig. 4) was a remarkable venture, bearing in mind that at that time experience with high-speed diesel engines was actually only with submarine engines having a maximum speed of not more than 500 rpm, whereas the new 150 hp Maybach diesel engine had an operating speed of 1300 rpm.

Solid fuel injection was then still at its beginning, so that the G 4 engine had air injection. Fig. 4 shows the relative sizes of the diesel engine and the engine-driven air compressor needed for air injection.

Opinion at that time - which many still hold today - was that the high-speed engines despite their obvious advantages with respect to weight and reduced space requirements would have a shorter service life, have less favorable consumption figures, and be more prone to trouble. This opinion chiefly resulted from the fact that many firms tried to increase the output of their existent low-speed engines merely by increasing the engine speed and neglecting any improvements in design. Naturally, this led to setbacks which in turn gave rise to such widespread opinion. The Maybach Motorenbau, which even then had many years of experience in the field of high-speed Otto engines, adopted new methods by designing a diesel engine especially for high speed and by aiming simultaneously at a corresponding increase of operating reliability and service life, two requirements obviously necessary for airship engines.

Since the first results with high-speed diesel engines were very encouraging, the development of these engines was continued and their output increased. The output of the first 6-cyl engine was soon raised to 210 hp at 1400 rpm. followed in 1930 by a 12-cyl V-engine providing 410 hp at 1400 rpm which, at a later date, was brought to 600 and 650 hp by exhaust gas supercharging (Fig. 5). In 1931 two of these 410 hp GO 5 engines were installed in the "Fliegender Hamburger," the first high-speed railcar-train of the German State Railway (Fig, 6). Its schedule speed of 77.6 mph, no less than the regular daily top speed of 100 mph, on the route between Hamburg and Berlin was a sensation at the time among railroad experts.

These Maybach engines of the GO-type - primarily designed as power units for railcars - set the fashion for diesel train operation of this power class in Europe until the year 1940. Hundreds of railcars and diesel trains were equipped with these engines in Germany, France, Belgium, Holland, Sweden, Norway, and Spain. Moreover, the engines met with favorable reception in naval construction, and were increasingly used for high-speed passenger boats, customs cruisers, yachts, coastguard crafts, and other craft.

The GO-type were designed so that the cylinder blocks of 6.30 in. bore and 7.88 in. stroke, made in one casting with the cylinder head, were assembled in an aluminum crankcase. The engines had direct injection, and four valves per cylinder controlled by overhead camshafts. The pistons were made of aluminum. Convenient mounting of running parts posed a special problem at the time for these powerful engines. White metal bearings could not be used, and lead-bronze bearings were still in too early stages of development. Therefore, it was decided to provide both the crankshaft and big ends with roller bearings, the composition of material and thermal treatment for which were the subject of lengthy and painstaking development by the company.

In operation, these engines attained mileages between general overhauls of as much as 100,000 miles (about 2500 operating hours) and more, which prior to World War I was regarded as quite satisfactory.

After termination of World War Il the situation was entirely different. Under the influence of the immense advance of the diesel locomotive in the United States and elsewhere, railway companies became increasingly interested in converting their heavy traction service to diesel operation.

If European manufacturers wanted to take part in this development, they had to counter the American diesel locomotive with its heavy and slow-running engine by a type of vehicle which would be available for a more universal application, offer more favorable weight conditions, and which in respect of output and service life would at least equal the heavy engines.

For Maybach Motorenbau, as the representative of the European trend, the demand for long life of the engine and its components became predominant. New lightweight small-size engines had to be designed which would be capable of offering mileages between overhauls at least equal to those then attained in rail traction by the finest types of slow-speed engines. Beyond that, these engines were supposed not only to compete successfully in the field of diesel traction, but also to conquer new fields of application.

Considering that for the aforementioned GO-type engines the roller bearings in the big end of the connecting rods were the limiting factor for the running performance, it was obvious that further development should aim at plain bearings for the big ends and a modernization of all running parts. The best solution in this direction appeared to be the discwebbed crankshaft, which inherently has greatly improved vibration characteristics, and which also provides ideal conditions for the entire bearing problem.

This was confirmed by the experience gained by Maybach Motorenbau during the last war with more than 50,000 Otto engines ranging from 300 to 700 hp at 3000 rpm, equipped with disc-webbed crankshafts and installed in heavy-duty military vehicles (tanks and heavy trucks). As a mater of fact, the adaptation of the disc-webbed crankshaft to the aforementioned pre-war GO-type, resulted in the change-over to the so-called GTO-"tunnel" type crankcase, representing a unique jump ahead in development.

Proof of this was the service results achieved by the German Federal Railway with its 600 bhp Maybach GTO type tunnel engines. It was revealed by spot checks that the bearings of these engines operating at 1400 rpm were still like new after distances of 300,000-400,000 miles. All other wearing parts showed such slight wear that a general overhaul was not anticipated before 600,000-700,000 miles, corresponding to about 12,000 to 15,000 operating hr. Early in 1955, these 600 bhp engines had exceeded an aggregate mileage of more than 6 million miles without any major engine overhauls (1).*

Today there are more than 1000 of these 12-cyl GTO engines rated up to 800 hp in service. Of these 840 are installed in shunting locomotives of the European standard type V.60, running for 20,000-25,000 hr before a piston check is scheduled.

Thus an optimum solution was found for the problems presented by the running gear of this engine type. For more powerful engines, however, with continuous outputs of about 100 hp per cylinder and with similar or rather better service behavior, one more step was necessary. That step was the introduction of the tunnel construction with roller main bearings surrounding the crank webs. The performance of this was proved first by thorough trials with powerful diesel engines of a speed range between 2400 and 2600 rpm. Eventually this led to the present Maybach MD tunnel engine, a design which has aroused great interest in the technical world. With this design, the problem of building a high-speed diesel engine with a running performance equal to, of better than that of a good slow-speed engine, can be regarded as solved.

Before giving a more detailed description of the MD design, it may be advantageous to discuss some general problems pertaining to high engine speed, and to report on the research techniques which contributed essentially to the development of the modern Maybach diesel engines.

General Questions Relating to High Engine Speed

It is, first of all, necessary to emphasize a fact often insufficiently realized by users: engine speed alone does not provide an adequate criterion for the genuine high-speed character of an engine. Even the mean piston speed does not provide such a criterion, since the mean effective pressure must also be taken into account, as well as whether the piston speed is used with high or low mean effective pressure at equal engine speed.

Some performance characteristics, important as far as engine loading is concerned, are now considered. Besides the absolute engine power N and the specific power per cylinder N/z, it is quite common to use the ratio of power per unit of swept volume N/V, characterizing the degree of utilization of the swept volume. The ratio of power per unit of piston area, N/F, also has a bearing in this connection: it signifies the total engine power related to the total area of the piston crowns. The piston crown is one of the engine parts subjected to the highest thermal load. To dissipate the heat accumulated here by means of conduction, radiation, or coolants is a matter of vital importance which may well be a limiting factor in increasing the specific engine power.

Strictly speaking, the ratios of power per unit of displaced volume, and power per unit of piston area can only provide a valid scale of comparison if the cylinders compared have the same size and shape. The influence of the stroke-to-bore ratio on these parameters has been investigated by Jaklitsch (2). Under certain limiting assumptions, he found that this influence is of the order of Vs/d, (s = stroke d =bore) a result derived statistically for diesel engines designed for aircraft and road vehicles, and it may be assumed, that this influence is equally applicable to larger types of diesel engines.

The fact that the ratio of power per unit of swept volume is influenced by the stroke-to-bore ratio can be explained by the fact that as this ratio changes there is also a change in the ratio of the heat-affected surface to the corresponding swept volume. With the longer stroke engine, where the piston diameter is smaller, the distance of heat flow from the center of the piston crown to the cylinder wall is smaller. Therefore, the temperatures at the center of the piston crown become higher if the piston diameter is increased. The results of similar investigations on the influence of cylinder bore and stroke-to-bore ratio on the piston temperature are plotted in Fig. 7. This graph shows that, even if mean effective pressure, piston speed, and stroke-to-bore ratio remain constant, an increase in the cylinder bore causes higher temperatures at the center of the piston crown. It will also be noted that the increase is steeper with short-stroke engines than with long-stroke engines. In order to obtain more realistic parameters for comparisons, the ratios of power per unit of swept volume and power per unit of piston area must be related to the stroke-to-bore ratio √s/d as follows:

"specific power per unit of swept volume" (N/V)$\sqrt{s/d}$

"specific power per unit of piston area" (N/F)$\sqrt{s/d}$

Apart from having such parameters as the specific power-to-volume ratio, and the specific power-to-piston area ratio, which are more indicative of the thermal loading of the pistons, it is also of importance to obtain criteria characterizing the mechanical stresses occurring in the rotating crankshaft assembly. This applies in particular to the big-end bearings, the loading of which might be characterized by the mass forces of the rotating and oscillating parts of the engine as well as by the maximum gas pressure. Assuming that, with the application of corresponding design principles, the weights of the rotating and oscillating engine parts vary with the third power of the piston diameter, a parameter a can be formed relating the mass forces to the unit of piston area, for example:

Jaklitsch calls this a value "running index" (2), and ascribes to it a major importance. The term "mass-force factor," however, appears to be more suitable as it relates more directly to the significance of the parameter. For it is accepted that, with a corresponding utilization of the cylinder diameter, the projected area of the big-end bearings increases with the square of the piston diameter, the massforce factor may well be accepted as a valid criterion for the bearing load due to the mass forces.

If the two engines compared are geometrically similar, that is, if the stroke-to-bore ratio remains the same, the above assumption is strictly true, as the masses in motion do vary with the third power of the piston diameter. In order to gain an idea in regard to the variation of the masses acting the big-end bearing When the stroke - to - bore ratio is modified, this ratio was altered in a V-engine and the variation of the masses acting on the bearing was calculated.

It was found that when the stroke-to-bore ratio was increased by as much as 60% the increase in weight was no more than 2-3%. It follows that, within reasonable limits, the mistake made by assuming that the variation of masses is proportional to d3 is insignificant.

F r o m this may be concluded that the engine with the most efficiently utilized crankshaft is the one which has the highest mass-force factor a, and, at the same time, can prove to be reliable in service and to have the service life expected for its application (5).

The diagram in Fig. 8 is intended to show the significance of the mass-force factor. For instance, if an engine has a speed of 1300 rpm and a cylinder bore of 7.28 in., with a stroke of 7.78 in. the mass-force factor becomes 672 ft/min?. If the same engine is accelerated to 1800 rpm, the mass-force factor increases to 1289 ft/min?. With a slow-speed engine of, say, 250 rpm, 17.72 in. bore and 25.58 in. stroke, the mass-force factor is no greater than 199 ft?/min?. Since both slow-speed and high-speed engines under consideration have actually been found to be very reliable in practice, it follows that the crankshaft assembly of the high-speed engine is much more efficiently utilized than that of the slow-speed engine.

In order to explain the significance of the mass-force factor still more in detail the characteristic parameters have been determined for some of the most important present-time locomotive engines of about 1000 bhp and more. The data are listed in Table 1. Four-cycle engines have been selected only so as to make a simple comparison of the parameters possible.

Two engines with equal characteristic parameters can be assumed to show equal service performance of their piston and crankshaft assemblies. This, however, holds only for engines of similar design. Therefore the parameters should never be compared on their own, but only under simultaneous consideration of the designs of the engines involved.

It appears from Table 1 that the specific power-to-piston area ratio of the locomotive engines - irrespective of their power and speed - is within the limits 210 to 345 bhp/ft?, excluding the engines represented by items 8 and 9, which will be discussed later. For the turbocharged stationary en-

gines, rated at continuous speeds from 250 to 600 rpm, the

respective p o w e r - t o - p i s t o n a r e a r a t i o s r a n g e a p p r o x i m a t e l y

between 93 and 186 bhp/ft?. From this comparison it be-

comes quite evident that the effort to ensure small bulk of

the locomotive engines to make them suitable for their ap-

plication, imposes high demands on engine design as well

as on material when operational reliability and life of the

wearing parts of both engine groups is to be the same. Even

more revealing are here the high mass-force factors for the

locomotive engines which range between 603 and 915 f t /

min', whereas the corresponding figures for stationary en-

gines lie only between 215 and 431 ft ⅔ /min?.

These simple considerations show that locomotive en-

gines which have proved satisfactory in practical operation

must be types of a highly meritorious design, for they have

withstood the imposition of quite severe demands.

In order to permit a comparison between different en-

gines on the basis of the criteria developed above, a com-

mon basis of "equal loading" has to be defined. "Equal

loading" of two engines specifically means that the mech-

anical and thermal loading of piston and crankshaft assem-

blies of the two engines is the same. This is characterized

by:

The conception of geometric similarity is understood to

cover not only the cylinders, (stroke and bore), but also all

those parts of the engine which are subjected to any stresses.

In all these comparisons, the mechanical efficiencies of the

engines are assumed to be constant.

From the well-known engine power equation:

Assuming that for two geometrically similar engines (s/d =

constant) the mean piston speed Cm the mean effective

pressure Po, the swept volume Ve and the ratio N/ are

constant, the correlation between power and number of cyl-

inders becomes:

Fig. 9 shows this functional relationship indicative of the

increase in output obtained with an engine of the same total

swept volume by increasing the number of cylinders.

For instance, if a plant hitherto driven by one slow-speed

12-cyl engine is driven, instead, by eight high-speed 12-

cyl engines which have altogether the same total swept vol-

ume as the one slow-speed engine, the output is doubled

although mechanical and thermal stresses in the piston and

crankshaft assembly are the same. If it were intended to

achieve this doubling of output by enlarging the swept vol-

ume of the single slow-speed engine, the swept volume of

that engine would have to be increased 2.8 times.

Finally, regarding the significance of the parameter of

specific power output per unit of piston area as a factor char-

acterizing engine loading, it may be used to divide engines

into different load or performance categories. For instance

certain ranges of this specific power-to-piston area factor

can be assigned to the category of engines of high specific

output, to the category of vehicle engines, the category of

marine engines, and so on. Under these assumptions can

be derived for N/F/Vs/d = constant and (s/d) = const.:

for a given load category and geometrically similar engines, the specific power-to-swept volume ratio is inversely pro-

portional to the bore.

Fig. 10 shows the relationship between the specific power-to-swept volume factor and the bore for different values of the specific power-to-piston area factor. For example, with geometrically similar engines of equal thermal and mechanical stressing of their piston resp, crankshaft assemblies, the doubling of the bore would halve the power per unit swept volume. In the same figure data are plotted for four high-speed engines belonging to four different performance categories, and one slow-speed marine propulsion engine.

The data of these engines are listed in Table 2.

Table 2 shows that a normal high-speed engine (for example, Engine No. 1) has the same power output per unit of piston area as the slow-speed marine engine No. 5, where the power of the latter must be regarded as being rather high compared with other similar engines. With high-speed engines, however, by means of suitable measures such as piston cooling the specific power output per unit of piston area can be increased for marine engines to 304 bhp/ft?, as shown for No. 3 engine. Yet this still does not exhaust the possibilities of increasing the specific power output. As No. 4 engine shows, by an increase in speed, mean effective pressure, and by other measures, the specific power-to-piston area factor of the same engine can be raised to beyond 600 bhp/ft?, a value which cannot be approached even remotely by slow-speed engines. This comparison clearly demonstrates the possibilities for high-speed engines to increase their specific power output; it also demonstrates the wide power range which can be obtained with one and the same engine.

The higher the specific power-to-piston area factor is of an engine which has proved to be reliable in actual service, the higher is the standard of technical development which the particular engine can be said to have reached. Hereby the measures are quite decisive by means of which higher specific power output per unit piston area has been attained, without exerting a detrimental influence on the reliability of the engine.

In the high-speed Maybach MD engines, a great step forward was made with the introduction of pressure-oil piston cooling . Fig. I l gives several temperatures measured at different power outputs in the piston crowns and piston ring lands of a pressure-oil cooled and a noncooled piston. The comparison clearly shows that although the power output per unit of piston area was considerably increased, the temperatures in the compression ring area, due to the effectiveness of the piston cooling, did not exceed those measured in the conventional standard design. These temperatures determine if seizing of the compression rings will occur. The only temperatures which increased slightly were those in the center of the piston crown. This is due to the considerably higher thermal loading and to a slightly smaller conductivity of the steel crown. Since, however, the crown of the cooled piston is of heat-resistant material and not of aluminum, this temperature rise has no detrimental effect on service life.

So far, in the comparative evaluation of different engines, the common base, namely the engine loading, was understood to refer exclusively to the piston and crankshaft assemblies. However, cylinder heads and liners are both subjected to high stresses. This is due to the explosion pressure and also to the heating of combustion chamber walls during the combustion process.

To simulate the maximum stresses in the cylinder head and liner due to ignition pressure Pz, the tangential stress