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High-Speed High-Output Diesel Engines

2026-10-04T14:32:33Z

Toro Andersen:

== 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)<math>\sqrt{s/d}</math>

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

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 engines, rated at continuous speeds from 250 to 600 rpm, the respective power - to - piston area ratios range approximately between 93 and 186 bhp/ft?. From this comparison it becomes quite evident that the effort to ensure small bulk of the locomotive engines to make them suitable for their application, 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 engines lie only between 215 and 431 ft ⅔ /min?.

These simple considerations show that locomotive engines 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 engines on the basis of the criteria developed above, a common basis of "equal loading" has to be defined. "Equal loading" of two engines specifically means that the mechanical and thermal loading of piston and crankshaft assemblies 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 cylinders 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 volume 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 volume 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 characterizing 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 proportional 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 °mech acting on the external diameter of a thick-walled cylinder may be taken. This is calculated by:

mech = Pz d2 / 2 (d + b) b

where w h e re .

d= Internal diameter

b = Wall thickness

Correspondingly, maximum tensile stress due to thermal load Otherm of the cylinder head and liner may be taken to be proportional to the tangential stress on the external diameter of a cylinder that is internally heated and externally cooled.

° therm = 1 - U 8(d + б ) I n (1 + 26)

where:

E = Modulus of elasticity

v= Poisson's number

B = Coefficient of expansion

AT = Difference between inner and outer wall temperature

It becomes apparent from Eg. 4 and 5 that where the same material is used, stressing of the walls in cylinder heads and liners depends on the ignition pressure Pz, on the difference in wall temperature ST, on the internal diameter d, and on the wall thickness &.

As regards ignition pressure Pz it may be assumed that this remains practically unchanged where engines of different size and uniform mean effective pressure Pe are concerned. However, when the latter decreases, the ignitionpressure will also decline.

Given the same material, the difference in wall temperatures AT is directly proportional to wall thickness and specific heat flow. From the known engine equations the following proportions will result:

N ~ Pe Cm d2 N~ Q

o ~ Pe Cm d2 power output per cylinder

total heat flow through the walls

forming the combustion space

specific heat flow through the walls

forming the combustion space

Thus the following proportion may be taken for the wall

temperature:

AT ~ Pe Cm S (6)

Assuming uniform piston speed Cm and uniform mean effective pressure Pe,

AT ~ 6 (7)

In Fig. 12 a computation diagram has been plotted from Egs. 4, 5, and 7. This diagram, though based on conditions pertaining to the Maybach MD diesel engine, permits assessing for other engines the relationship between wall thickness of cylinder heads or liners, and mechanical and thermal stressing of walls at constant mean effective pressure. In this connection it should be noted that for the Maybach MD engine, of 7.38 in. bore and an average wall thickness of the cylinder head of 0.34 in., the mechanical and thermal stress figures have been entered as 100%.

Fig. 12 clearly shows that in a slow-running engine with

cylinders twice the diameter of the MD engine, with the

same mean effective pressure, and the same cylinder head

and liner wall thickness, the mechanical stresses in cylinder

head and liner walls are twice as high as in the case of the

quick-running engine. Thermal stressing of the walls only

differs very slightly. If, in this comparison, mechanical

stressing is kept constant in the slow-running engine, then

doubling the wall thickness will be required, in which case

thermal load will increase 2.4 times compared to the high-

speed engine.

For constant mechanical and thermal wall stressing and

equal piston speeds, a correlation between cylinder bore and

mean effective pressure was calculated from the above pro-

portions and plotted in Fig. 13. This shows clearly the ex-

tent to which the high-speed engine excels the medium- or

slow-speed engine in regard to the mean effective pressure

that can be realized. As an example, a diesel engine with

twice the cylinder diameter of the Maybach MD engine, and

having the same mean piston speed and the same mechan-

ical and thermal loadings of cylinder head and liner walls,

can be operated only at two-thirds the mean effective pres-

sure of the corresponding MD engine. This fact should be

taken into account because the possible thermal and me-

chanical loading of cylinder head and liner materials is prac-

tically the decisive factor today in limiting the specific power

output of a diesel engine.

For a long time, there was a widespread idea that high-

speed engines, because of their higher rpm and their gen-

erally higher piston speed have a greater liner wear than low-

speed engines. This idea has been discarded even by the

supporters of the slow-running engine.

It is argued today that the large bore of a slow-speed en-

gine also permits substantial overall wear before the cylin-

221

der liner must be replaced. Departing from the assumption

that the absolute wear per 1000 operating hr is about equal

in either case, the conclusion is that the cylinder liners of

a high-speed engine should be replaced at considerably shorter

i n t e r v a l s . thus resulting, as a rule, in higher costs for

spare parts, because any wear phenomena at this spot would

have an approximately constant relationship to the other

wearing parts of an engine (6). This conclusion, however,

stems from an assumption which in no way is represented by

actual existing conditions.

When studying technical literature, time and again one

finds the indication of 0.004 in. per 1000 per operating hr

as an average figure in regard to liner wear of large slow-

running marine engines. Depending on the nature of the fuel

used, this figure varies slightly upwards or downwards.

the other hand, there are operational results available today

from high-speed engines in railroad service which permit

certain conclusions. Fig. 14 shows the wearing quality of

some typical high-speed engines selected from a larger num-

ber of engines of the same type which for more than 5 yr

have been operated by the German Federal Railway (7).

engines concerned are 12-cyl Maybach engines, type GTO,

developing 600 bhp at 1400 rpm with a bore of 6.30 in. and

a stroke of 7.88 in., this being equal to a mean effective

pressure of 113.5 psi and a piston speed of 32.8 fps. These

engines still have uncooled pistons and therefore may read-

ily be compared with high-speed engines of another type.

With an average wear of 0.0063 in. related to 7500 hr, cyl-

inder wear in these engines is approximately 0.0008 in. for

every 1000 operating hr. These liner wear figures were all

measured where wear is greatest, namely, at the top dead

center of the topmost compression ring. As in the case of

cylinder bore, the permissible overall wear is in the neigh-

borhood of 0.016 in., reboring in connection with a major engine overhaul will not be required before some 20,000 op-

erating hr.

In this respect the objection may be raised that contrary

to the propulsion of sea-going ships, engines operated in rail

traction are not continously running at full load. This ob-

jection cannot be denied. However, it is a well-known fact

that continuous load changes, as they are the rule in rail

traction, subject the engine to increased thermal stressing,

and dust content of the induced air is another drawback for

railway engines. Wear figures obtained for more recently

installed MD-engines will be discussed later.

Apart from the wear resistance of an engine, the load on

the crank assembly is another factor of importance regarding engine service life, and here something of the influence

of increased engine speed may be shown. Characteristics

representing constant load on the main bearings of the crank-

shaft of an MD engine, the load being caused by ignition pres-

sure and mass forces, are plotted in Fig. 15 as a function of

torque and rotational speed. This graph shows that the load

on the main bearings at constant engine torque increases with

decreasing engine speed. Therefore, operating the engine

along a constant torque characteristic causes considerable

additional stressing of the main bearings. In order to obtain

power outputs as high as possible, it is from this point of

view more advisable to increase the speed rather than operate

at low speed and high torque. Operating the engine along the

propeller characteristic appears therefore to yield best serv-

ice results (8).

Based on the resulting bearing load the service life of the

roller main bearings was calculated. The lines of constant

service life have been plotted in Fig. 16 as a function of

power output and engine speed. It should be noted that this

calculation only applies to crankshafts with roller bearings.

From this graph can be seen that regarding service life con-

ditions are most favorable when with decreasing power out-

put engine speed is not reduced. It will also be noted that

if an engine is operated along the propeller curve, at a load

of about 83% of continuous power output, bearing life is

doubled, whereas with constant torque operation,e v e n a t

the lowest engine speeds there will be no increase in service

life. This again shows that increasing engine speed rather

than torque gives best results with respect to service life of

the crankshaft roller bearings.

The influence of high-speed on weight can readily be

seen in Table 1. When considering the column of weights

in this table the engines may be classified in three groups.

The first group with a weight-to-power ratio of about 20

to 22 1b/hp comprises the engines indicated under Nos. 1 and

These are 4-stroke engines with speeds up to 850 rpm,

built by European makers. Their specific power-to-piston area ratios and their mass-force factors are comparatively

low. The engines are quite heavy, thus having an essen-

tial bearing on locomotive weight. In the case of electric

transmissions the low speed will furthermore result in high

generator weights, unless a more expensive type incorporat-

ing two crankshafts and a step-up gear transmission for the

generator drive is adopted as for engine No. 2.

The second group refers to engines with a uniform weight

of some 16-18 1b/hp at 1000 rpm. Typical representatives

of this group are the two engine types nos. 3 and 4; these

are American 4-stroke locomotive engines (Alco and Cooper

Bessemer) of identical cylinder dimensions which are widely

used. Note that the weight reduction of these engines is

linked with an increase of the specific power-to-piston area

factor to about 344 bhp/ft?. For engines which have proved

to be satisfactory in actual service such values bear testi-

mony of a quite advanced stage in development. For com-

pleteness sake it should be mentioned, that into this cate-

gory belongs a widely used 2-stroke engine (General Motors)

16-cyl, weighing 15.8 tons, developing 1950 bhp at 835 rpm

corresponding to a weight-to-power ratio of 16.3 1b/bhp.

A definite step forward in regard to weight was the in-

troduction of engines of group III - engine nos. 5, 6 and 7 -

having a weight-to-power ratio of 6.5 to 1 1b/bhp.

These engines rotate at 1500 rpm and must therefore be classified

as high-speed engines. They originated in Europe and some

of them have gone through many years of development and

constitute an advanced form of powerful railcar engine.

In contrast to the United States, diesel rail traction in

Europe until World War Il was exclusively by the diesel rail

car. Here reference is made to the numerous high-speed

rail-cars and streamlined trains equipped with high-speed

Maybach diesel engines. The former German State Railways and other railroad companies operated such trains even

before the last war in multiple-units up to five cars powered by up to 2000 hp. This experience, supplemented by knowl-

edge obtained during the war years in all fields of engine

construction enabled the industry, by introducing engines

of this group, to set entirely new standards for the construc-

tion of powerful locomotives with a favorable power-to-

weight ratio.

Though it has not yet been generally recognized today,

these engines have no higher specific loading in regard to

their power ratings than typical low-speed engines, since

their specific power-to-piston area ratio, ranging from 278

to 334 bhp/ft?, is even lower than the corresponding values

of the American 4stroke engines mentioned before.

The mass-force factor of these engines, ranging from 790

to 915 ft/min?, is slightly higher; but this has not given

rise to any difficulties in practice, since the use of better

materials and improved machining techniques increased the

load bearing capability of the parts - especially of the con-

necting rod. Economically the use of better material was

possible as the bulk of the parts decreased.

Table 1 gives evidence of the substantial advantages of-

fered by high-speed engines in regard to weight. This re-

duced weight of the high-speed engine is an absolute ne-

cessity for the construction of powerful diesel locomotives,

and becomes even more important if for economic reasons

locomotives with not more than four axles are considered.

=== Development of Measuring Techniques for High-Speed Engines ===
Formerly, in development and testing of internal com-

bustion engines one had to rely fully on results obtained by

test bed endurance trials and on ensuing experience gained

i n p r a c t i c a l o p e r a t i o n . With this kind of testing it could

often take a long time before any weak spots in the engines

c a m e t o l i g h t . However, especially in the last 10 yr, elec-

trical measuring techniques have been so improved that it

is now possible to actually measure the mechanical load to

which an internal combustion engine is subjected. In the

case of high-speed engines the requirements with respect to

measuring methods are considerablym o r e s u n d e n o w i n ?

to the higher frequencies, so that recent development of

measuring techniques is of particular benefit to the high-

speed engines.

In order to anticipate the reliability of individual ma-

chine members, it is of prime importance to know the fa-

tigue strength of the material. The results from fatigue

strength tests can be presented in the form of a diagram in

which the number of load alternations prior to the fatigue

fracture is entered as a function of the ultimate strength.

was recognized by Wohler as long as a hundred years ago that

there is an interrelationship between the fatigue strength of a

steel and the load cycle, and that a material which has with-

stood a certain load level for a definite number of alternat-

ing stress cycles will not break. As can be seen from Fig.

17 this load level does not alter even when the number of

load cycles is increased from 10 million to 200 million (Fig.

18). For both high-speed and low-speed engines the number of 10 million load alternations is attained relatively quickly,

so that for the design the fatigue strength and not the tensile

strength has to be taken into account. It is quite natural that

when investigating the fatigue strength of a material there

will be some dispersion owing to irregularities in the surface

structure, and it so happens that in endurance tests not a

single fatigue characteristic is obtained but rather a range,

of which the lower stray field limitation supplies the prac-

tical values for calculating the operating reliability. In such

investigation regard must be paid to whether there are pure

tension-compression or bending stresses, or if there is a non-

uniform stress distribution. In the case of applying a bend-

ing load on small diameter rods or rods with notches, higher

alternating stress amplitudes might be acceptable, because

residual stresses as a result of machining the surface layer

may become particularly perceptible here.

In the Maybach stress laboratory various grades of steel

have been examined in respect of their endurance strengths.

In Fig. 19 the bending fatigue strength of various steels has

been plotted as a function of the tensile strength. The in-

dividual values clearly indicate that the service life of an-

nealed (hardened and tempered) steels increases with the

tensile strength. Therefore, the alloying constituents are

only of importance inasmuch as they ensure perfect and

thorough hardening. As a comparison with these measure-

ments, a characteristic curve "b" by Wellinger and Gimmel

(9) has been added to Fig. 19, showing general agreement

with the findings of Maybach Motorenbau as represented by

curve "a."

Apart from the influence of tensile strength on fatigue

strength, exhaustive studies were also made regarding the

influence exercised by surface finish and material purity.

Some results of such an investigation are shown in Fig. 20.

The graph shows that machining may have a considerable influence on the fatigue strength of the material, and that

under certain circumstances the advantage of high strength

as offered by high tensile steels may be offset entirely

by their poor machinability. As indicated in the graph, the

decrease in fatigue strength in the case of a forged surface

may be quite considerable. Similarly unfavorable condi-

tions are to be expected where impurities and heat-treat-

ment defects are present. However, by the use of modern

test methods it is possible today to ascertain a desired de-

gree of surface finish and purity of a material so that, when

properly inspected, the minimum fatigue strength of the lat-

ter can be realized. Apart from the surface finish, the in-

fluence exercised by fretting corrosion on fatigue strength

is also of considerable importance.

In this connection the bending fatigue strength of a high

tensile steel clamped between two jaws of unhardened car-

bon steel was determined by the Maybach stress laboratory.

It is shown in Fig. 21 that due to the action of fretting cor-

rosion phenomena the bending fatigue strength of this rod

was reduced to as low as 28,400 psi. Another test speci-

men of the same material, soft-nitrided, attained a bend-

ing fatigue strength of 71,000 psi under the same circum-

s t a n c e s . Consequently, the elimination of fretting corrosion

by nitriding and the compressive stresses thereby induced in the surface, increased the bending fatigue strength of this

material by 150%.

Once the fatigue strength of a material is known, the de-

termination of the actual loads on individual structural mem-

bers is necessary, and particularly the pattern of the stresses

induced by the loads. An excellent way to ascertain the

direction of major deformation and/or principal stresses, even

in the case of structural members with complicated contours,

is the brittle lacquer method which, as long ago as 1924,

was evolved by Maybach Motorenbau in connection with the

development of engines for the airship "ZR III," the "Los

Angeles," which on Oct. 13, 1924, started for its epic flight

across the Atlantic Ocean.

In this method, the component considered is coated with

a special lacquer, in the surface of which cracks - the so-

called elongation lines - will show up at stresses below the

elastic limit of the structural members. These elongation

lines, originating always vertically to the tensile load, first

appear in the surface area which is subjected to highest

stresses. As an example, Fig. 2 shows the image of elongation

lines on a flat bar under a bending load. In the area under

tensile load (bottom of Fig. 22) the elongation lines are di-

rected perpendicular to the tensile strains, and in the area

of compressive load (top of Fig. 22) they are directed in line

with the compressive strain. When the stress pattern has been

determined, a suitable strain-measuring instrument may be

used to measure the actual magnitude of the stresses.

Practical application may be shown by the images of

elongation lines in a fork connecting rod in Fig. 23. In the

rod on the left hand side of the figure the lacquer was ap-

plied prior to applying the tensile load, and on the right

hand side prior to relieving the load. In this way, the first

and second directions of main stresses are indicated.

subsequent strain measuring with strain gages pasted perpen-

dicularly to the elongation lines, the stress distribution in

the connecting rod can be determined. Local stress peaks

thus found can be eliminated or reduced to acceptable lim-

its by suitable modifications in design.

Due to the smallness of present-day strain gages it is easy

to measure the stresses even in lubricating oil holes having,

for example, a diameter of only 0.315 in. The effects of

such holes in crankshafts have been determined by the May-

bach stress laboratory (10). As may be seen from Fig. 24,

stress concentrations are best removed by widening the hole

in a slightly conical form. In spite of this, however, round-

ing off the rim of the hole will still be necessary to obtain a faultless surface. Care must be taken to ensure that the

location of the hole subjected to maximum stress is not on

the outside of the rim, but somewhat more to the inside.

Thanks to the high standard of modern measuring tech-

niques it is also possible to determine the stresses in the

crankshafts of engines in motion. In this respect, investi-

gations were made at Friedrichshafen on a 16-cyl MD en-

gine. Strain gages were applied to each web of the crank-

shaft, and the leads were brought to the outside through slip

rings. By way of an automatically operating multiple switch

gear these gages were connected with an oscilloscope, which

filmed the load diagrams of the individual strain gages in

rapid succession. At the same time, by means of a torsional

vibration gage developed by the Works, torsional vibration

amplitides were measured at the free end of the engine crank-

shaft and likewise filmed on an oscilloscope.

The load diagrams established at the crank webs are so-

called "cumulative curves" of the local crankshaft loading.

These curves were evaluated under two aspects: (1) the oc-

currence of maximum loading, (maximum crankshaft torque

in regard to magnitude and location), and (2) by harmonic

analysis, in order to obtain a comparison between the mea-

sured local load and the results of the torsional vibration computation supplemented by the measured vibration am-

plitude at the free end of the shaft.

In Fig. 25 two series of typical load diagrams are shown

for the eight throws of a 16-cyl V-engine crankshaft (crank-

pin 1 on the driving end). During this measuring procedure

the engine output was 2500 bhp at 1800 rpm. The diagrams

to the left of the schematic shaft were plotted for the shaft

not fitted with a torsional vibration damper, whereas those

directly to the right of the crankshaft were plotted for an

assembly fitted with a very effective type of viscous damper.

The test record sbove each series of load diagrams repre-

sents the corresponding torsional vibration amplitude mea-

sured at the free end of the crankshaft. When examining

the illustrations it must be borne in mind that owing to vary-

ing degrees of sensitivity of the strain gages, the scale for

the amplitude in the load diagrams is not the same for all

t r o w s . The diagrams at the extreme right represent the

summation of the crank torques as calculated from the in-

dicator diagrams of individual cylinders. The scale of the

amplitude of the calculated torque diagrams always coin-

cides with that of the corresponding measured-load diagrams,

whereas the scale of the abscissae is considerably compressed.

It appears from the comparison that the measured torques

in the throws, whether or not a damper is fitted, are actually

smaller than those computed. In addition, there is evidence

that the stresses when a damper is used are smaller still;

and the shape of the load diagrams for the shaft having an

efficient damper will thus be approximate to the shape of the

corresponding calculated diagrams.

In order to give an idea of the stress distribution along

the shaft, the maximum positive and negative stresses se-

lected from the corresponding load diagrams (Fig. 26) have

been plotted against the length of the shaft for operation

both with and without a torsional vibration damper. It will be noted that maximum torque occurs in the middle of the

shaft. For comparative purposes, the calculated torques

have been shown as well.

From this investigation one may appreciate that with an

installation having favorable torsional vibration character-

istics, the installation of a damper may bring little further

reduction in the stresses beyond those shown in Fig. 26: on

the other hand, lack of a damper in an installation with un-

favorable characteristics may bring higher stresses than those

indicated in the figure.

The amplitudes of the torsional vibrations at the free end,

and the load diagrams measured at the crank-throws, were

analyzed electronically in order to check the procedure by

w h i c h t h e s t r e s s e s d u e t o t o r s i o n a l v i b r a t i o n s c a n b e e s t i -

mated. Operation without a damper resulted in greatest

loads occurring close to the third natural frequency; for this

resonance range the measured and computed load ampli-

tudes of the harmonics of higher orders are plotted in Fig.

27 against the shaft length. Here harmonics of the 4th and

6.5th orders showed actual values up to 145% above those

calculated, whereas for all other orders the values deter-

mined were lower than those calculated. It must be borne

in mind that the correspondingly analyzed components of

the forces arising from gas pressure and inertia are not con-

tained in the calculated amplitudes of the harmonic oscil-

lations, but are included in the analyzed amplitudes actually

measured. As regards order of magnitude, it may be ap-

preciated from these measurements, an approximate deter-

mination of the stresses in the crankshaft can be obtained

by the vibration calculation for the case of resonance, sup-

plemented by the measurements of the vibration amplitude

at the free end of the shaft.

An electric indicator of Maybach's own design with a

seismic system is used to measure torsional vibrations. It

consists of a heavy symmetrical mass which revolves to gether with the engine crankshaft. The mass itself is con-

nected to the crankshaft by means of four spiral springs and

a damping device. Whenever there is any torsional vibration

of the crankshaft, the mass will maintain its uniform rotary

motion and the crankshaft will thus move against the mass.

This movement of the crankshaft causes a self-inductance

variation in coils forming a bridge fed with alternating cur-

rent. Supply of feeding voltage and pick-up of measuring

voltage takes place through a slip ring transmitter. The tor-

sional vibration indicator is operated over-critically. By

adequately tuning both mass and springs its natural frequency

could be brought down to 5 cps.

It is an advantage of this type of torsion indicator that

observation is possible during the test, so that increase in

vibration limits can be noted at once, and the test values

recorded remotely.

Pressures in the cylinders of high-speed engines are pref-

erably measured today by indicators based on the piezo-

electric principle. Very exact measurements are possible

with such indicators, provided that the length of the hole

connecting cylinder and indicator is shorter than its diame-

ter. Should this not be the case the test results will be grossly

falsified due to induced compression waves in the con-

necting channel. Fig. 28 shows a typical test record of an

8-cyl Maybach MD 440 engine operating at 1000 bhp at

1500 rpm where five items are recorded simultaneously: noz-

zle lift in the unit injector, plunger stroke in the unit in-

jector, pressure distribution in the precombustion chamber,

pressure distribution in the main cylinder space, and crank

angle marking. Valve lift and stroke of the plunger were

measured directly in the unit injector by inductive means,

the two pressure diagrams with the aid of piezoelectric quartz

indicators installed in the cylinder head, and the crank-an-

gle marking was obtained by means of make-and-break con-

r a c i s . All five test results were transmitted to a reed-type

oscillograph where they were recorded photographically. The diagram of the nozzle valve lift in Fig. 28 clearly

reveals with the marked kink the actual commencement of

injection (8.5 deg btdc). Regarding pressure distribution in

the precombustion chamber, a rapid rise approximately 4

deg crank angle after opening of the nozzle valve can be

recognized. This 4 deg crank angle, corresponding to 0.5

millisec, represents the ignition delay of the fuel.

When comparing the pressure distribution in the precom-

bustion chamber with that in the cylinder it will be readily

observed that during the process of combustion there is com-

pression-wave action between precombustion chamber and

cylinder. This, however, is less perceptible in the cylinder

than in the precombustion chamber.

A further example of measuring instruments which have

been employed for the progressive development of Maybach

engines is the air-speed indicator of the firm's own design.

This device can be arranged immediately before the inlet

valve in the intake pipe and is used to measure the speed

and the direction of air flowing through the inlet valve. In

this way, it is possible to supplement the pressure measure-

ments in cylinder, exhaust and intake pipes carried out to

determine optimum scavenging conditions. In principle, the

air speed indicator consists of a small baffle plate which, by

a torsion rod, is connected with two small soft-iron plates;

these in turn are arranged in the air gap of inductive coils.

Under the action of the air flow impact on the baffle, the

small soft-iron plates are displaced, whereby self-induct-

ance variations are set up in the coils forming the bridge cir-

cuit of a carrier frequency amplifier. The self-inductance

variations (proportional to the variation in air speed) are

transformed by the amplifier into changes in voltage which

are recorded by an oscilloscope. The natural frequency of

the air speed indicator is above 100 cps. Damping of the

coil system is achieved by filling the latter with a viscous

damping compound.

=== General Description of MD-Engine Range ===
Having dealt with the history of Maybach Motorenbau and with the techniques available today for continuing research now given on and development of these engines, certain deails are now given on the progressive development of the MD-engines.

The experience gained with high-speed diesel engines since 1923, and particularly tests with powerful diesel engines in the speed range of 2400-2600 rpm, eventually led to the Maybach tunnel engine of the MD type running generally at a speed around 1500 rpm. Like earlier Maybach engines, the MD tunnel-type engines are single-acting 4-stroke engines. Ranging from 4- to 16-cyl, they are built either as in-line engines or as V models, naturally-aspirated or turbocharged. Outstanding design features of the MD engines are:

1. Forged one-piece disc-webbed crankshaft carried in a single piece tunnel housing.

2. Pressure-oil cooled aluminum pistons (forced circulation cooling system) with detachable steel crowns.

3. Individual cylinder heads with central pre-combustion chamber, and six small well-cooled valves, with automatic backlash adjustment.

4. Unit injectors for each cylinder.

5. Tunnel crankcase fabricated of steel castings, steel plates and profiles.

6. Large number of repetitive components for the whole engine range.

These points will now be discussed more in detail.

Disc-Webbed Crankshaft and Tunnel-Form Crankcase - In the MD-design, almost the entire length of the engine is used for main and big-end bearings; the crankwebs, which in a conventional engine must be regarded as a loss in this

respect, are here encircled by roller main bearings.

results in the following advantages:

Short engine length with minimum distances between cyl-

inders - this is of great advantage regarding elastic defor-

mation and vibration problems.

Considerable reduction of bending stresses to which the

crankshaft is subjected, due to the crankwebs forming the

Particularly good conditions for all crankshaft bearings

as these are generously dimensioned.

Wide connecting-tod bearings of width nearly equal to

diameter are used without danger of edge loading. There-

fore liquid film lubrication can be assured under all oper-

Due to the comparatively small dimensions of such a disc-

webbed crankshaft, questions of material, flow of fibers,

surface hardness and quality are easier to solve than they

are for larger-sized crankshafts, especially of slow-running

engines.

As a result of its rugged construction and particularly of

its shorter length a disc-webbed crankshaft is at least 40%

stiffer than a conventional crankshaft having a pin of the

s a m e diameter. In order to demonstrate this effect more

clearly the natural frequencies of a conventional crankshaft

and a disc-webbed shaft have been compared in Fig. 29.

The two shafts are illustrated in Fig. 30. The conventional

crankshaft of the Maybach GO engine has a stiffness between

two adjacent piston masses of c = 18.2 × 106 in-lb/rad. Thus the natural frequency of the engine alone, with flywheel

but without vibration damper, becomes n,= 5870 min"!.

Within the speed range between 500 and 1400 rpm the 6th

order which is very strong is located at 980 rpm. On the

other hand, the disc-webbed crankshaft of the rebuilt GTO

MD engine has a stiffness of c = 53.2 × 106 in.-lb/rad

and a natural frequency of the engine alone without vibra-

tion damper of ny = 10,550 min"!.

With this high first nat-

ural frequency the 6th order lies outside the working speed

range, and since higher orders do not need to be taken into

consideration, the whole range from 500 to 1400 rpm is free

from resonance.

The advantage of the disc-webbed crankshaft design (11)

is not as easy to see for torsional loads as for bending loads,

because transfer of force from the crankpin to the web can-

not be avoided even in a disc-webbed shaft. Nevertheless,

the disc-webbed crankshaft still has superior form factors Ok

related to torsional loads: ax = 2.32 for the disc-webbed

crankshaft as compared with 0x = 2.79 for the conventional

crankshaft (Fig. 33).

Form factor refers to the ratio between the stresses actually

measured in the fillet and those calculated for the pin.

[[Category:Articles]]

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Worlds largest engine

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Toro Andersen:

== What is the World's Largest Engine? ==
The phrases '''"world’s largest engine"''' and '''"world’s biggest engine"''' are inherently ambiguous because they lack a clear frame of reference. "Biggest" and "largest" can mean entirely different things depending on context, industry, and measurement criteria. To define the ultimate winner, one must first establish the metric of comparison.

=== Sector Differences ===
Different industries define an "engine" differently based on their primary function:

* A massive '''marine diesel engine''' powers global container ships, built for immense continuous torque at low revolutions.
* A '''gas turbine''' like the GE9X drives commercial wide-body aircraft, focusing on high bypass airflow and efficiency.
* A '''rocket engine''' such as the Apollo-era ''F-1'' or the SpaceX Raptor produces unmatched raw thrust relative to its own physical footprint.

Each of these machines can justifiably claim the title of "world’s biggest engine" within its respective sector.

=== Core Metrics for Comparison ===
Without specifying "biggest by what metric," the statement remains vague. The primary technical metrics used to rank large engines include:

# '''Physical Size (Mass and Dimensions):''' Total weight (measured in metric tonnes) and external footprint (height, length, width). This is the key metric for stationary powerplants and marine vessels where installation space and structural mass dictate engineering limits.
# '''Power Output (Horsepower or Kilowatts):''' How much work the engine performs per unit of time (kW or bhp).
# '''Thrust (Newton or Pound-force):''' The critical metric for jet and rocket propulsion, evaluating forward force via high-velocity propellant ejection (kN or MN).
# '''Hybrid Metrics:''' Engine torque (rotational force, kNm) or power-to-weight ratio (hp/kg), which provide insights into mobile efficiency and performance applications.

=== The Element of Time ===
The target shifts constantly across history. What was considered a monumental engine in 1895 or 1920 is a medium-sized auxiliary unit by modern industrial standards. Therefore, an explicit timeframe is required. For example, if looking for the engine with the '''absolute largest cylinder bore''' ever put into production, that honor historically belongs to the '''Fiat GMT 1060S''', boasting a colossal bore diameter of '''1,060 mm (1.06 meters)'''.

=== Hybrid Metrics (Torque or Power‑to‑Weight) ===
Alternative measures such as torque (rotation force), or power‑to‑weight ratio (hp / kg) can provide more context, especially in performance or mobile applications like vehicles or aircraft.

=== The Historical Contenders ===
When evaluating physical scale and displacement, the conversation shifts exclusively to '''low-speed, two-stroke crosshead diesel engines'''. The most famous names and licensors through history include:

* '''Burmeister & Wain (B&W):''' Danish pioneers of marine diesel technology, later merged into MAN B&W. Renowned for long-stroke efficiency.
* '''Doxford:''' British engineering firm famous for its distinctive "Opposed Piston" (OP) design, offering exceptional mechanical balance.
* '''Götaverken:''' Swedish shipyard and engine designer that produced highly regarded uniflow-scavenged two-stroke main engines.
* '''Fiat GMT:''' Grandi Motori Trieste (Italy), which pushed the extreme physical boundaries of cylinder diameter during the 1970s.
* '''MAN (Maschinenfabrik Augsburg-Nürnberg):''' German industrial heavyweight. Through MAN Energy Solutions, they currently design a dominant share of the world's large commercial marine engines.
* '''Mitsubishi UEC:''' The only surviving entirely Japanese-designed large two-stroke marine engine series, known for highly compact footprints.
* '''Stork-Werkspoor:''' Dutch manufacturer celebrated for robust historical configurations like the mid-century ''HOTLo'' series.
* '''Sulzer:''' Swiss brand (later Wärtsilä, now WinGD), designer of the reigning physical record holder for the world's largest operational engine line, the ''RT-flex96C''

== Big engines, Large engines. ==
Engines, large ones, mostly mounted in ships, and there has been many famous names up thru the years.

What was a big engine in 1895 is shure a small engine today.

If we want the engine with the biggest bore, that honor probaly goes to the 1060 mm bore Fiat GMT

== Modern Giants (Year 2000 and Onwards) ==
The following table provides the comprehensive technical specifications of the largest low-speed, two-stroke crosshead engines developed in the 21st century. These power the modern fleet of mega-container vessels.
{| class="wikitable"
|+
! colspan="18" |Slow Speed 2 stroke crosshead engines from 2000 and onwards - The BIG ones
|-
!Engine
type / serie
!Variant
!Cylinders
!Bore
mm
!Stroke
mm
!Volume
pr cyl
liter
!Volume
engine
liter
!kW
!bhp
!BHP pr
cyl
!Torque
kNm
!RPM
!Length
mm
!Width
mm
!Height
mm
!Weight
tonnes
!Year in
service
!Notes
|-
|Sulzer RT-flex 96c
|14
|14
|960
|2500
|1809
|25334
|80080
|108920
|7780
|7498
|102
|27313
|4480
|13519
|2300
|2004
|
|-
|MAN B&W K108
|MC-C
|14
|1080
|2660
|2432
|34048
|97300
|132328
|9450
|9880
|94
|28500
|4900
|15200
|
|(2004)
|Planned, not made
|-
|MAN B&W G95
|C9
|12
|950
|'''3460'''
|'''2470'''
|'''29640'''
|82440
|112126
|'''9340'''
|9840
|80
|22000
|5150
|15000
|
|
|
|-
|MAN B&W K98
|ME7 / MC-C7
|14
|'''980'''
|2660
|2007
|22638
|'''87220'''
|'''119307'''
|8521
|8590
|97
|27885
|4370
|13450
|2405
|2005
|
|-
|MAN B&W S90ME
|C 9.2
|14
|900
|3260
|2040
|28560
|81340
|110630
|7902
|
|84
|
|
|
|
|
|
|}

== Historical Engine Catalog (20th Century Milestones) ==
This reference index lists the technical layouts, physical boundaries, and design strategies deployed by global engine developers throughout the 1900s.
<blockquote>Note: Some of these data are sourced via unreliable sources, this message stays until we have confirmed them. </blockquote>
{| class="wikitable"
|+Older big engines - Major engine makers with "large" engines made
!Engine Builder / Licensor
!Trunk Cross head
!2/4 Stroke
!Engine type
!Single Double Oposing
!Larges cylinder config
!Bore mm
!Stroke mm
!Volume pr cylinder
!BHP pr cyl
!MEP bar
!RPM
!Piston speed m/sec
!Weight tonn
!Year in service
!Historical Significance / Notes
|-
|Alfred Holt
|CH
|2
|Scott-Still (Holt Type)
|DA
|4
|560
|915
|225.3
|625
|5.1
|120
|3.66
|210
|1923
|Early highly complex custom fleet deployment of the Scott-Still hybrid oil/steam concept for Blue Funnel Line.
|-
|AEG
|CH
|2
|Standard Series
|DA
|6
|700
|1200
|461
|1250
|4,8
|90
|3,6
|550
|1928
|An early pioneer in double-acting two-stroke designs aimed at maximizing volumetric power.
|-
|AEG-Hesselman
|CH
|2
|Hesselman Series
|SA
|4
|600
|2320
|656.0
|750
|4.9
|80
|3.09
|320
|1928
|Pioneered Hesselman's proprietary mechanical airless fuel injection system on large two-stroke layouts.
|-
|[[Burmeister & Wain]]
|CH
|2
|K98FF
|SA
|12
|980
|2000
|1508.59
|3800
|10,8
|103
|6,86
|1380
|1968
|The world's first true "super-bore" production engine, breaching the 3,500 hp/cylinder mark.
|-
|Deutsche Werft
|CH
|2
|DW-AEG Geared Series
|SA
|6
|700
|1200
|461.8
|1000
|4.5
|110
|4.40
|420
|1930
|Built primarily under AEG/Hesselman design frameworks for standard large cargo vessel configurations.
|-
|[[Doxford]]
|CH
|2
|J Type
|OP
|6
|760
|580+1600 2180
|988
|2537
|9,5
|119
|6,13
|580
|1959
|Famous British opposed-piston design utilizing a balanced three-crankpin assembly per cylinder.
|-
|[[Fiat GMT]]
|CH
|2
|1060S
|SA
|12
|'''1060'''
|1900
|1676.7
|4000
|10,0
|106
|6,71
|1640
|1971
|King of the Bore. Retains the record for the largest cylinder diameter ever put into serial production.
|-
|Fullagar (Cammel Laird)
|CH
|2
|Fulgar Series
|OP
|6
|584
|1829
|490
|450
|5,2
|90
|5,48
|240
|1924
|Unique compact opposed-piston concept featuring oblique tie-rods linking upper and lower crossheads.
|-
|Franco Tosi Meccanica
|CH
|2
|Type D
|SA
|10
|770
|1500
|698.4
|750
|5.2
|105
|5.25
|580
|1933
|Italian high-power marine platform heavily inspired by early loop-scavenged crosshead designs.
|-
|[[Götaverken]]
|CH
|2
|DM 850/1700
|SA Uniflow
|10
|850
|1700
|964
|2100
|9,8
|115
|6,52
|800
|1960
|Proprietary Swedish uniflow-scavenged design utilizing distinct exhaust valves driven by cam-segments on the crankshaft.
|-
|Harland & Wolff
|CH
|2
|H&W-B&W 840 Series
|DA
|10
|840
|1500
|831.3
|2000
|6.5
|115
|5.75
|980
|1930
|Massive double-acting units powering high-profile ocean liners like the MV Britannic.
|-
|Ingersoll-Rand
|Trunk
|4
|Price System (PR)
|SA
|8
|559
|610
|149.7
|150
|6.2
|250
|5.08
|85
|1924
|Prominent early American heavy trunk diesel utilizing a specialized dual-opposed fuel injection design.
|-
|Krupp
|CH
|2
|Krupp-Diesel Series
|SA
|8
|750
|1500
|662.7
|700
|4.8
|95
|4.75
|480
|1926
|Early heavy-displacement marine designs, utilizing port scavenging and robust cast structural components.
|-
|McIntosh and Seymour
|CH
|4
|Standard Crosshead
|SA
|6
|813
|1524
|791.2
|650
|5.4
|95
|4.83
|510
|1925
|Large-bore four-stroke design widely selected for early US Shipping Board motor conversion programs.
|-
|[[MAN]]
|CH
|2
|KSZ105/180
|SA
|12
|1050
|1800
|1558.62
|4000
|10,7
|108
|6,48
|1520
|1969
|Developed during the "super-bore" era to directly rival Sulzer’s and Fiat’s massive configurations.
|-

|[[Mitsubishi Heavy Industries|Mitsubishi]] (Kobe)
|CH
|2
|MS Series
|SA
|9
|750
|1500
|662.7
|1200
|5.1
|110
|5.50
|540
|1932
|Mainstay Japanese domestic marine design, evolving directly from early technical assistance agreements.
|-
|[[Nobel]]
|CH
|2
|Nobel-Diesel Marine
|SA
|6
|650
|1200
|398.2
|400
|4.6
|105
|4.20
|220
|1922
|Developed by the Nobel family’s specialized engineering works, advancing early open-crankcase marine diesels.
|-
|North British Diesel Engine Company (NBDEC)
|CH
|2
|Sliding Cylinder
|DA
|6
|673
|1194
|424.7
|500
|4.9
|96
|3.82
|390
|1924
|Highly unconventional layout where the entire cylinder moved vertically to uncover scavenging ports.
|-
|Nordberg
|CH
|2
|RF-90
|SA
|12
|889
|1752
|1087.4
|710
|5.8
|120
|7.01
|920
|1946
|Notable large-scale American-built engine line deployed widely across municipal powerplants and heavy cargo ships.
|-
|Polar
|Trunk
|2
|Polar T-Series
|SA
|6
|500
|700
|137.4
|400
|5.2
|200
|4.66
|115
|1935
|Atlas-Polar design showcasing highly efficient port-scavenging loops on a compact trunk frame.
|-
|Richardsons Westgarth
|CH
|2
|RW
|DA
|6
|699
|1200
|460
|1050
|4,7
|105
|4,2
|410
|1926
|British-designed, fully fabricated all-welded structural engine concept that was highly innovative for its era.
|-
|[[Sulzer]]
|CH
|2
|RND 105
|SA
|12
|1050
|1800
|1558.62
|4000
|10,5
|108
|6,48
|1540
|1967
|Pioneered pure loop-scavenging at an enormous scale, avoiding cylinder-head exhaust valves for structural simplicity.
|-
|Stork
|CH
|2
|HOTLo 85/170
|SA
|9
|850
|1700
|964
|1850
|7,5
|115
|6,52
|680
|1954
|Noted for its uniflow execution using four distinct, smaller exhaust valves clustered around a single cylinder head.
|-
|[[Stork-Werkspoor]]
|CH
|2
|SW 280 Series
|SA
|9
|850
|1700
|964.4
|2250
|9.5
|115
|6.52
|720
|1970
|Later unified technical iteration combining Werkspoor's engineering lines into Stork's uniflow platforms.
|-
|Still & Straight Diesel Scott Engines
|CH
|2
|Scott-Still
|SA
|6
|686
|1118
|413
|415
|6,1
|116
|4,32
|310
|1934
|A highly advanced hybrid layout using exhaust waste heat to boil steam, which mechanically drove the underside of the piston.
|-
|Swan Hunter
|CH
|2
|Type B
|SA
|6
|600
|1200
|339.3
|350
|4.5
|100
|4.00
|190
|1922
|Development of AB Diesels engines utilizing early British collaborative structural components.
|-
|UEC (Mitshubishi)
|CH
|2
|UEC 85/180D
|SA
|12
|850
|1800
|1021
|2300
|9,2
|115
|6,9
|920
|1970
|Early premier domestic Japanese uniflow-scavenged design utilizing a unique multi-valve configuration.
|-
|Vickers
|CH
|4
|Vickers Commercial
|SA
|8
|762
|1143
|521.2
|450
|5.2
|110
|4.19
|340
|1921
|Early solid-injection champion, eliminating high-pressure blast air compressors for extreme reliability.
|-
|Werkspoor
|CH
|4
|Lugt Engine (or Lugt SA)
|SA
|6
|680
|1250
|453.9
|600
|5.5
|115
|4.79
|290
|1948
|Post-war innovative four-stroke design featuring a lightweight, valveless cylinder cover and turbocharging.
|-
|Worthington
|CH
|2
|Worthington Double-Acting
|DA
|4
|711
|965
|383.1
|725
|5.0
|95
|3.06
|260
|1926
|Notable American foray into double-acting two-stroke operations aimed at US maritime shipping lines.
|}

Worlds largest engine

2026-10-02T13:38:08Z

Toro Andersen: /* Historical Engine Catalog (20th Century Milestones) */

== What is the World's Largest Engine? ==
The phrases '''"world’s largest engine"''' and '''"world’s biggest engine"''' are inherently ambiguous because they lack a clear frame of reference. "Biggest" and "largest" can mean entirely different things depending on context, industry, and measurement criteria. To define the ultimate winner, one must first establish the metric of comparison.

=== Sector Differences ===
Different industries define an "engine" differently based on their primary function:

* A massive '''marine diesel engine''' powers global container ships, built for immense continuous torque at low revolutions.
* A '''gas turbine''' like the GE9X drives commercial wide-body aircraft, focusing on high bypass airflow and efficiency.
* A '''rocket engine''' such as the Apollo-era ''F-1'' or the SpaceX Raptor produces unmatched raw thrust relative to its own physical footprint.

Each of these machines can justifiably claim the title of "world’s biggest engine" within its respective sector.

=== Core Metrics for Comparison ===
Without specifying "biggest by what metric," the statement remains vague. The primary technical metrics used to rank large engines include:

# '''Physical Size (Mass and Dimensions):''' Total weight (measured in metric tonnes) and external footprint (height, length, width). This is the key metric for stationary powerplants and marine vessels where installation space and structural mass dictate engineering limits.
# '''Power Output (Horsepower or Kilowatts):''' How much work the engine performs per unit of time (kW or bhp).
# '''Thrust (Newton or Pound-force):''' The critical metric for jet and rocket propulsion, evaluating forward force via high-velocity propellant ejection (kN or MN).
# '''Hybrid Metrics:''' Engine torque (rotational force, kNm) or power-to-weight ratio (hp/kg), which provide insights into mobile efficiency and performance applications.

=== The Element of Time ===
The target shifts constantly across history. What was considered a monumental engine in 1895 or 1920 is a medium-sized auxiliary unit by modern industrial standards. Therefore, an explicit timeframe is required. For example, if looking for the engine with the '''absolute largest cylinder bore''' ever put into production, that honor historically belongs to the '''Fiat GMT 1060S''', boasting a colossal bore diameter of '''1,060 mm (1.06 meters)'''.

=== Hybrid Metrics (Torque or Power‑to‑Weight) ===
Alternative measures such as torque (rotation force), or power‑to‑weight ratio (hp / kg) can provide more context, especially in performance or mobile applications like vehicles or aircraft.

=== The Historical Contenders ===
When evaluating physical scale and displacement, the conversation shifts exclusively to '''low-speed, two-stroke crosshead diesel engines'''. The most famous names and licensors through history include:

* '''Burmeister & Wain (B&W):''' Danish pioneers of marine diesel technology, later merged into MAN B&W. Renowned for long-stroke efficiency.
* '''Doxford:''' British engineering firm famous for its distinctive "Opposed Piston" (OP) design, offering exceptional mechanical balance.
* '''Götaverken:''' Swedish shipyard and engine designer that produced highly regarded uniflow-scavenged two-stroke main engines.
* '''Fiat GMT:''' Grandi Motori Trieste (Italy), which pushed the extreme physical boundaries of cylinder diameter during the 1970s.
* '''MAN (Maschinenfabrik Augsburg-Nürnberg):''' German industrial heavyweight. Through MAN Energy Solutions, they currently design a dominant share of the world's large commercial marine engines.
* '''Mitsubishi UEC:''' The only surviving entirely Japanese-designed large two-stroke marine engine series, known for highly compact footprints.
* '''Stork-Werkspoor:''' Dutch manufacturer celebrated for robust historical configurations like the mid-century ''HOTLo'' series.
* '''Sulzer:''' Swiss brand (later Wärtsilä, now WinGD), designer of the reigning physical record holder for the world's largest operational engine line, the ''RT-flex96C''

== Big engines, Large engines. ==
Engines, large ones, mostly mounted in ships, and there has been many famous names up thru the years.

What was a big engine in 1895 is shure a small engine today.

If we want the engine with the biggest bore, that honor probaly goes to the 1060 mm bore Fiat GMT

== Modern Giants (Year 2000 and Onwards) ==
The following table provides the comprehensive technical specifications of the largest low-speed, two-stroke crosshead engines developed in the 21st century. These power the modern fleet of mega-container vessels.
{| class="wikitable"
|+
! colspan="18" |Slow Speed 2 stroke crosshead engines from 2000 and onwards - The BIG ones
|-
!Engine
type / serie
!Variant
!Cylinders
!Bore
mm
!Stroke
mm
!Volume
pr cyl
liter
!Volume
engine
liter
!kW
!bhp
!BHP pr
cyl
!Torque
kNm
!RPM
!Length
mm
!Width
mm
!Height
mm
!Weight
tonnes
!Year in
service
!Notes
|-
|Sulzer RT-flex 96c
|14
|14
|960
|2500
|1809
|25334
|80080
|108920
|7780
|7498
|102
|27313
|4480
|13519
|2300
|2004
|
|-
|MAN B&W K108
|MC-C
|14
|1080
|2660
|2432
|34048
|97300
|132328
|9450
|9880
|94
|28500
|4900
|15200
|
|(2004)
|Planned, not made
|-
|MAN B&W G95
|C9
|12
|950
|'''3460'''
|'''2470'''
|'''29640'''
|82440
|112126
|'''9340'''
|9840
|80
|22000
|5150
|15000
|
|
|
|-
|MAN B&W K98
|ME7 / MC-C7
|14
|'''980'''
|2660
|2007
|22638
|'''87220'''
|'''119307'''
|8521
|8590
|97
|27885
|4370
|13450
|2405
|2005
|
|-
|MAN B&W S90ME
|C 9.2
|14
|900
|3260
|2040
|28560
|81340
|110630
|7902
|
|84
|
|
|
|
|
|
|}

== Historical Engine Catalog (20th Century Milestones) ==
This reference index lists the technical layouts, physical boundaries, and design strategies deployed by global engine developers throughout the 1900s.
{| class="wikitable"
|+Older big engines - Major engine makers with "large" engines made
!Engine Builder / Licensor
!Trunk
Cross head
!2/4
Stroke
!Engine type
!Single
Double

Oposing
!Larges
cylinder

config
!Bore
mm
!Stroke
mm
!Volume
pr cylinder
!BHP pr
cyl
!MEP
bar
!RPM
!Piston
speed

m/sec
!Weight
tonn
!Year in
service
!Historical Significance / Notes
|-
|Alfred Holt
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|AEG
|CH
|2
|Standard Series
|DA
|6
|700
|1200
|461
|1250
|4,8
|90
|3,6
|550
|1928
|An early pioneer in double-acting two-stroke designs aimed at maximizing volumetric power.
|-
|AEG-Hesselman
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|[[Burmeister & Wain]]
|CH
|2
|K98FF
|SA
|12
|980
|2000
|1508.59
|3800
|10,8
|103
|6,86
|1380
|1968
|The world's first true "super-bore" production engine, breaching the 3,500 hp/cylinder mark.
|-
|Deutsche Werft
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|[[Doxford]]
|CH
|2
|J Type
|OP
|6
|760
|580+1600
2180
|988
|2537
|9,5
|119
|6,13
|580
|1959
|
|-
|[[Fiat GMT]]
|CH
|2
|1060S
|SA
|12
|'''1060'''
|1900
|1676.7
|4000
|10,0
|106
|6,71
|1640
|1971
|King of the Bore. Retains the record for the largest cylinder diameter ever put into serial production.
|-
|Fullagar (Cammel Laird)
|CH
|2
|Fulgar Series
|OP
|6
|584
|1829
|490
|450
|5,2
|90
|5,48
|240
|1924
|Unique compact opposed-piston concept featuring oblique tie-rods linking upper and lower crossheads.
|-
|Franco Tosi Meccanica
|
|2
|Type D
|SA
|10
|770
|
|
|750
|
|
|
|
|1933
|
|-
|[[Götaverken]]
|CH
|2
|DM 850/1700
|SA Uniflow
|10
|850
|1700
|964
|2100
|9,8
|115
|6,52
|800
|1960
|Proprietary Swedish uniflow-scavenged design utilizing distinct exhaust valves driven by cam-segments on the crankshaft.
|-
|Harland & Wolff
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|Ingersoll-Rand
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|Krupp
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|McIntosh and Seymour
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|[[MAN]]
|CH
|2
|KSZ105/180
|SA
|12
|1050
|1800
|1558.62
|4000
|10,7
|108
|6,48
|1520
|1969
|Developed during the "super-bore" era to directly rival Sulzer’s and Fiat’s massive configurations.
|-
|[[Mitsubishi Heavy Industries|Mitsubishi]] (Kobe)
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|[[Nobel]]
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|North British Diesel Engine Company (NBDEC)
|
|
|
|SA
|
|673
|1194
|
|
|
|96
|
|
|
|
|-
|Nordberg
|CH
|2
|RF-90
|
|
|889
|1752
|
|
|
|
|
|
|
|Notable large-scale American-built engine line deployed widely across municipal powerplants and heavy cargo ships.
|-
|Polar
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|Richardsons Westgarth
|CH
|2
|RW
|DA
|6
|699
|1200
|460
|1050
|4,7
|105
|4,2
|410
|1926
|British-designed, fully fabricated all-welded structural engine concept that was highly innovative for its era.
|-
|[[Sulzer]]
|CH
|2
|RND 105
|SA
|12
|1050
|1800
|1558.62
|4000
|10,5
|108
|6,48
|1540
|1967
|Pioneered pure loop-scavenging at an enormous scale, avoiding cylinder-head exhaust valves for structural simplicity.
|-
|Stork
|CH
|2
|HOTLo 85/170
|SA
|9
|850
|1700
|964
|1850
|7,5
|115
|6,52
|680
|1954
|Noted for its uniflow execution using four distinct, smaller exhaust valves clustered around a single cylinder head.
|-
|[[Stork-Werkspoor]]
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|Still & Straight Diesel
Scott Engines
|CH
|2
|Scott-Still
|SA
|6
|686
|1118
|413
|415
|6,1
|116
|4,32
|310
|1934
|A highly advanced hybrid layout using exhaust waste heat to boil steam, which mechanically drove the underside of the piston.
|-
|Swan Hunter
|
|2
|Type B
|
|
|
|
|
|
|
|
|
|
|1922
|Development of AB Diesels engines
|-
|UEC (Mitshubishi)
|CH
|2
|UEC 85/180D
|SA
|12
|850
|1800
|1021
|2300
|9,2
|115
|6,9
|920
|1970
|Early premier domestic Japanese uniflow-scavenged design utilizing a unique multi-valve configuration.
|-
|Vickers
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|Werkspoor
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|Worthington
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|}

== Historical Engine Catalog (20th Century Milestones) ==
This reference index lists the technical layouts, physical boundaries, and design strategies deployed by global engine developers throughout the 1900s.
{| class="wikitable"
|+Older big engines - Major engine makers with "large" engines made
!Engine Builder / Licensor
!Trunk Cross head
!2/4 Stroke
!Engine type
!Single Double Oposing
!Larges cylinder config
!Bore mm
!Stroke mm
!Volume pr cylinder
!BHP pr cyl
!MEP bar
!RPM
!Piston speed m/sec
!Weight tonn
!Year in service
!Historical Significance / Notes
|-
|Alfred Holt
|CH
|2
|Scott-Still (Holt Type)
|DA
|4
|560
|915
|225.3
|625
|5.1
|120
|3.66
|210
|1923
|Early highly complex custom fleet deployment of the Scott-Still hybrid oil/steam concept for Blue Funnel Line.
|-
|AEG
|CH
|2
|Standard Series
|DA
|6
|700
|1200
|461
|1250
|4,8
|90
|3,6
|550
|1928
|An early pioneer in double-acting two-stroke designs aimed at maximizing volumetric power.
|-
|AEG-Hesselman
|CH
|2
|Hesselman Series
|SA
|4
|600
|2320
|656.0
|750
|4.9
|80
|3.09
|320
|1928
|Pioneered Hesselman's proprietary mechanical airless fuel injection system on large two-stroke layouts.
|-
|[[Burmeister & Wain]]
|CH
|2
|K98FF
|SA
|12
|980
|2000
|1508.59
|3800
|10,8
|103
|6,86
|1380
|1968
|The world's first true "super-bore" production engine, breaching the 3,500 hp/cylinder mark.
|-
|Deutsche Werft
|CH
|2
|DW-AEG Geared Series
|SA
|6
|700
|1200
|461.8
|1000
|4.5
|110
|4.40
|420
|1930
|Built primarily under AEG/Hesselman design frameworks for standard large cargo vessel configurations.
|-
|[[Doxford]]
|CH
|2
|J Type
|OP
|6
|760
|580+1600 2180
|988
|2537
|9,5
|119
|6,13
|580
|1959
|Famous British opposed-piston design utilizing a balanced three-crankpin assembly per cylinder.
|-
|[[Fiat GMT]]
|CH
|2
|1060S
|SA
|12
|'''1060'''
|1900
|1676.7
|4000
|10,0
|106
|6,71
|1640
|1971
|King of the Bore. Retains the record for the largest cylinder diameter ever put into serial production.
|-
|Fullagar (Cammel Laird)
|CH
|2
|Fulgar Series
|OP
|6
|584
|1829
|490
|450
|5,2
|90
|5,48
|240
|1924
|Unique compact opposed-piston concept featuring oblique tie-rods linking upper and lower crossheads.
|-
|Franco Tosi Meccanica
|CH
|2
|Type D
|SA
|10
|770
|1500
|698.4
|750
|5.2
|105
|5.25
|580
|1933
|Italian high-power marine platform heavily inspired by early loop-scavenged crosshead designs.
|-
|[[Götaverken]]
|CH
|2
|DM 850/1700
|SA Uniflow
|10
|850
|1700
|964
|2100
|9,8
|115
|6,52
|800
|1960
|Proprietary Swedish uniflow-scavenged design utilizing distinct exhaust valves driven by cam-segments on the crankshaft.
|-
|Harland & Wolff
|CH
|2
|H&W-B&W 840 Series
|DA
|10
|840
|1500
|831.3
|2000
|6.5
|115
|5.75
|980
|1930
|Massive double-acting units powering high-profile ocean liners like the MV Britannic.
|-
|Ingersoll-Rand
|Trunk
|4
|Price System (PR)
|SA
|8
|559
|610
|149.7
|150
|6.2
|250
|5.08
|85
|1924
|Prominent early American heavy trunk diesel utilizing a specialized dual-opposed fuel injection design.
|-
|Krupp
|CH
|2
|Krupp-Diesel Series
|SA
|8
|750
|1500
|662.7
|700
|4.8
|95
|4.75
|480
|1926
|Early heavy-displacement marine designs, utilizing port scavenging and robust cast structural components.
|-
|McIntosh and Seymour
|CH
|4
|Standard Crosshead
|SA
|6
|813
|1524
|791.2
|650
|5.4
|95
|4.83
|510
|1925
|Large-bore four-stroke design widely selected for early US Shipping Board motor conversion programs.
|-
|[[MAN]]
|CH
|2
|KSZ105/180
|SA
|12
|1050
|1800
|1558.62
|4000
|10,7
|108
|6,48
|1520
|1969
|Developed during the "super-bore" era to directly rival Sulzer’s and Fiat’s massive configurations.
|-

|[[Mitsubishi Heavy Industries|Mitsubishi]] (Kobe)
|CH
|2
|MS Series
|SA
|9
|750
|1500
|662.7
|1200
|5.1
|110
|5.50
|540
|1932
|Mainstay Japanese domestic marine design, evolving directly from early technical assistance agreements.
|-
|[[Nobel]]
|CH
|2
|Nobel-Diesel Marine
|SA
|6
|650
|1200
|398.2
|400
|4.6
|105
|4.20
|220
|1922
|Developed by the Nobel family’s specialized engineering works, advancing early open-crankcase marine diesels.
|-
|North British Diesel Engine Company (NBDEC)
|CH
|2
|Sliding Cylinder
|DA
|6
|673
|1194
|424.7
|500
|4.9
|96
|3.82
|390
|1924
|Highly unconventional layout where the entire cylinder moved vertically to uncover scavenging ports.
|-
|Nordberg
|CH
|2
|RF-90
|SA
|12
|889
|1752
|1087.4
|710
|5.8
|120
|7.01
|920
|1946
|Notable large-scale American-built engine line deployed widely across municipal powerplants and heavy cargo ships.
|-
|Polar
|Trunk
|2
|Polar T-Series
|SA
|6
|500
|700
|137.4
|400
|5.2
|200
|4.66
|115
|1935
|Atlas-Polar design showcasing highly efficient port-scavenging loops on a compact trunk frame.
|-
|Richardsons Westgarth
|CH
|2
|RW
|DA
|6
|699
|1200
|460
|1050
|4,7
|105
|4,2
|410
|1926
|British-designed, fully fabricated all-welded structural engine concept that was highly innovative for its era.
|-
|[[Sulzer]]
|CH
|2
|RND 105
|SA
|12
|1050
|1800
|1558.62
|4000
|10,5
|108
|6,48
|1540
|1967
|Pioneered pure loop-scavenging at an enormous scale, avoiding cylinder-head exhaust valves for structural simplicity.
|-
|Stork
|CH
|2
|HOTLo 85/170
|SA
|9
|850
|1700
|964
|1850
|7,5
|115
|6,52
|680
|1954
|Noted for its uniflow execution using four distinct, smaller exhaust valves clustered around a single cylinder head.
|-
|[[Stork-Werkspoor]]
|CH
|2
|SW 280 Series
|SA
|9
|850
|1700
|964.4
|2250
|9.5
|115
|6.52
|720
|1970
|Later unified technical iteration combining Werkspoor's engineering lines into Stork's uniflow platforms.
|-
|Still & Straight Diesel Scott Engines
|CH
|2
|Scott-Still
|SA
|6
|686
|1118
|413
|415
|6,1
|116
|4,32
|310
|1934
|A highly advanced hybrid layout using exhaust waste heat to boil steam, which mechanically drove the underside of the piston.
|-
|Swan Hunter
|CH
|2
|Type B
|SA
|6
|600
|1200
|339.3
|350
|4.5
|100
|4.00
|190
|1922
|Development of AB Diesels engines utilizing early British collaborative structural components.
|-
|UEC (Mitshubishi)
|CH
|2
|UEC 85/180D
|SA
|12
|850
|1800
|1021
|2300
|9,2
|115
|6,9
|920
|1970
|Early premier domestic Japanese uniflow-scavenged design utilizing a unique multi-valve configuration.
|-
|Vickers
|CH
|4
|Vickers Commercial
|SA
|8
|762
|1143
|521.2
|450
|5.2
|110
|4.19
|340
|1921
|Early solid-injection champion, eliminating high-pressure blast air compressors for extreme reliability.
|-
|Werkspoor
|CH
|4
|Lugt Engine (or Lugt SA)
|SA
|6
|680
|1250
|453.9
|600
|5.5
|115
|4.79
|290
|1948
|Post-war innovative four-stroke design featuring a lightweight, valveless cylinder cover and turbocharging.
|-
|Worthington
|CH
|2
|Worthington Double-Acting
|DA
|4
|711
|965
|383.1
|725
|5.0
|95
|3.06
|260
|1926
|Notable American foray into double-acting two-stroke operations aimed at US maritime shipping lines.
|}

Worlds largest engine

2026-10-02T13:36:34Z

Toro Andersen: /* Historical Engine Catalog (20th Century Milestones) */

Worlds largest engine

2026-10-02T13:31:29Z

Toro Andersen: /* Historical Engine Catalog (20th Century Milestones) */

Worlds largest engine

2026-10-02T13:20:17Z

Toro Andersen: /* Modern Giants (Year 2000 and Onwards) */

Worlds largest engine

2026-10-02T13:19:52Z

Toro Andersen:

Worlds largest engine

2026-09-30T17:02:10Z

Toro Andersen:

== What is the worlds largest engine? The worlds biggest engine? ==
The phrases "world’s largest engine" and "world’s biggest engine" is ambiguous because it lacks a clear reference frame or category; “biggest” and "largest" can mean several different things depending on context, purpose, and measurement criteria. So to clearly define it, one needs to agree on what it is biggest or largest in.

== Sector Differences ==
Different industries define “engine” differently. A massive marine diesel engine powers container ships, a jet turbine like the GE9X drives aircraft, and rocket engines such as the Raptor or F‑1 produce unmatched thrust. Each could justifiably claim the title of “world’s biggest engine” within its sector, leading to natural ambiguity.

== Comparative Language Issues ==
The term also suffers from linguistic ambiguity — "biggest" and "largest" are a relative adjective that implies comparison but not a fixed technical measure. Without specifying “biggest by what metric” or “in which category,” the statement becomes vague and open to misinterpretation.

In short, “world’s biggest engine” is ambiguous because “biggest” can signify entirely different metrics — mass, volume, or output — and “engine” itself applies to many unrelated machines across industries, each optimized for different functions.

== Time issue ==
As time progresses, the target shifts. What was the biggest engine in 1920 is by any measure a medium size engine at todays standards. So we need to issue a timeframe to the equation too.

== Physical Size (Mass and Dimensions) ==
If “biggest” refers to physical scale, then we measure total weight and external dimensions (height, length, width). This matters for stationary powerplants and ship engines, where installation space and total mass are significant.

== Power Output (Horsepower or Kilowatts) ==
“Biggest” sometimes refers to power output — how much work the engine performs per unit time, measured in horsepower (hp) or kilowatts (kW). Proton rocket engines, jet turbines, and electrical generators are often ranked by this value.

== Thrust (Newton or Pound‑force) ==
For jet and rocket engines, thrust is the key metric — the total forward force produced by propellant ejection. Rocket classifications like F‑1 or Raptor engines use kilonewtons (kN) or meganewtons (MN) to define “bigness”

== Hybrid Metrics (Torque or Power‑to‑Weight) ==
Alternative measures such as torque (rotation force), or power‑to‑weight ratio (hp / kg) can provide more context, especially in performance or mobile applications like vehicles or aircraft.

== Big engines, Large engines. ==
Engines, large ones, mostly mounted in ships, and there has been many famous names up thru the years.

What was a big engine in 1895 is shure a small engine today.

If we want the engine with the biggest bore, that honor probaly goes to the 1060 mm bore Fiat GMT

== The Contenders - short, brief history. ==

It must be a crosshead engine, two stroke.

Burmeister & Wain

Doxford

Götaverken

Fiat GMT

MAN

Mitsubishi 12UEC85/180D

Stork-Werkspoor

Sulzer

{| class="wikitable"
|+
! colspan="18" |Slow Speed 2 stroke crosshead engines from 2000 and onwards - The BIG ones
|-
!Engine
type / serie
!Variant
!Cylinders
!Bore
mm
!Stroke
mm
!Volume
pr cyl
liter
!Volume
engine
liter
!kW
!bhp
!BHP pr
cyl
!Torque
kNm
!RPM
!Length
mm
!Width
mm
!Height
mm
!Weight
tonnes
!Year in
service
!Notes
|-
|Sulzer RT-flex
|14
|14
|960
|2500
|1809
|25334
|80080
|108920
|7780
|7498
|102
|27313
|4480
|13519
|2300
|2004
|
|-
|MAN B&W K108
|MC-C
|14
|1080
|2660
|2432
|34048
|97300
|132328
|9450
|
|94
|
|
|
|
|2004
|Planned, not made
|-
|MAN B&W G95
|C9
|12
|950
|'''3460'''
|'''2470'''
|'''29640'''
|82440
|112126
|'''9340'''
|
|80
|22000
|
|15000
|
|
|
|-
|MAN B&W K98
|ME7 / MC-C7
|14
|'''980'''
|2660
|1617
|22638
|'''87220'''
|'''119307'''
|8521
|
|97
|27885
|4370
|13450
|2405
|2005
|
|-
|MAN B&W S90ME
|C 9.2
|14
|900
|3260
|2040
|28560
|81340
|110630
|7902
|
|84
|
|
|
|
|
|
|}
{| class="wikitable"
|+Older big engines - Major engine makers with "large" engines made
!Engine maker
!Trunk
Cross head
!2/4
Stroke
!Engine type
!Single
Double

Oposing
!Larges
cylinder
config
!Bore
mm
!Stroke
mm
!Volume
pr cylinder
!BHP pr
cyl
!MEP
bar
!RPM
!Piston
speed
m/sec
!Weight
tonn
!Year in
service
|-
|Alfred Holt
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|AEG
|
|
|
|
|
|700
|1200
|
|
|
|90
|
|
|1928
|-
|AEG-Hesselman
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|[[Burmeister & Wain]]
|CH
|
|K98FF
|SA
|
|980
|2000
|1508.59
|3800
|10,8
|103
|6,86
|
|1968
|-
|Deutsche Werft
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|[[Doxford]]
|CH
|
|J Type
|OP
|
|760
|520+1600
|
|2537
|
|119
|
|
|1959
|-
|[[Fiat GMT]]
|CH
|
|1060S
|
|
|'''1060'''
|1900
|1676.7
|4000
|10,0
|106
|6,71
|1640
|1971
|-
|Fullagar
|
|2
|
|OP
|
|584
|1829
|
|
|
|90
|
|
|1924
|-
|Franco Tosi Meccanica
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|[[Götaverken]]
|
|2
|850/1700
|
|10
|850
|1700
|
|
|
|
|
|
|
|-
|Harland & Wolff
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|Ingersoll-Rand
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|Krupp
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|McIntosh and Seymour
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|[[MAN]]
|CH
|
|KSZ105/180
|
|
|1050
|1800
|1558.62
|4000
|10,7
|108
|6,48
|
|1969
|-
|[[Mitsubishi Heavy Industries|Mitsubishi]] (Kobe)
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|[[Nobel]]
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|North British Diesel Engine Company (NBDEC)
|
|
|
|SA
|
|673
|1194
|
|
|
|96
|
|
|
|-
|Nordberg
|
|
|RF90
|
|
|889
|1752
|
|
|
|
|
|
|
|-
|Polar
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|Richardsons Westgarth
|
|2
|
|DA
|
|699
|1200
|
|
|
|105
|
|
|
|-
|[[Sulzer]]
|CH
|
|RND 105
|
|
|1050
|1800
|1558.62
|4000
|10,5
|108
|6,48
|
|1967
|-
|Stork
|
|
|HOTLo
|
|
|850
|1700
|
|
|
|
|
|
|
|-
|[[Stork-Werkspoor]]
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|Still & Straight Diesel / Scott Engines
|
|2
|
|SA
|
|686
|1118
|
|
|
|116
|
|
|1934
|-
|Swan Hunter
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|UEC (Mitshubishi)
|
|
|
|
|9
|800
|3040
|
|5470
|
|
|
|
|
|-
|Vickers
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|Werkspoor
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|-
|Worthington
|
|
|
|
|
|
|
|
|
|
|
|
|
|
|}

Digby B. Morton

2026-09-01T20:01:42Z

Toro Andersen:

Morton patented a opposed piston engine in the 1950s.

He was situated in Red Hill, Papakura, in New Zealand.

[[File:1957 Digby Morton.png|thumb|1954 - Engine]]---

== Patent application ==
Jan. 10, 1956

Digby B. Morton, Digby B. Morton, Jr., and Graeme E. Morton, Kilbirnie, Wellington, New Zealand Application April 20, 1953, Serial No. 349,834

Claims priority, application New Zealand May Z, 1952 7 Claims. (Cl. 123-51) This invention relates to reciprocating internal combustion engines of the opposed piston type wherein the power is transmitted from each opposed piston through a kinematic chain to a driven member or members, such, for example, as a common rotating crank shaft; and is particularly though not solely of use in two cycle compression ignition engines of this type.

It is known that internal combustion engines of the type described possess thermodynamic advantages, particularly in the case of two cycle compression ignition engines of this type, but in present engines of the type described, disadvantages exist in that the kinematic chains connecting the pistons to the driven member or members have disadvantages, for example being unduly complicated or expensive to manufacture. Furthermore, such engines are not compact and must be built larger and heavier than would otherwise be necessary in order to accommodate the kinematic chains.

lt is therefore an object of the present invention to provide an internal combustion engine and preferably a two cycle compression ignition engine, of the type described, which is simple in construction and therefore cheap to manufacture, and which is compact in size compared with present engines of this type.

It is a further object of the present invention to provide an internal combustion engine of the type described which may be more easily dismantled and reassembled than present engines of this type.

lt is a still further object of the present invention in a preferred form to provide an internal combustion engine in which scavenging or charging air may be provided in sufficient quantity to give eicient scavenging and charging in a simple and convenient manner.

Accordingly in one aspect the invention consists in an internal combustion engine of the type described, wherein the kinematic chain connecting each piston to the driven member or members includes an oscillating lever which is so slidably associated with said piston that said lever will oscillate when said piston reciprocates.

In a further aspect the invention consists in an internal combustion engine of the type described, wherein the kinematic chain connecting each piston to the driven member or members includes an oscillating lever, which is so associated with said piston intermediate of the length of the piston, that said lever will oscillate when said piston reciprocates.

An advantage of the invention in its preferred form resides in the automatic allowance for misalignment between the cylinder and the crankshaft by the means used to connect the pistons, the gudgeon pins and the rocker arms. Y

One preferred form of the invention will now be described with reference to the accompanying drawings in which:

Figure l is a vertical cross-section on the line 1 1, Figure 2, of an engine constructed according to the invention, Y n

Figure 2 is a vertical cross-section on the line 2-2,

Figure l, the connecting rods being omitted and some parts being shown in elevation,

Figure 13 is a section of a piston and its associated oscillating lever on the lines 3 3, Figure l,

Figure 4 is a fragmentary side view, partly in broken lines, showing the securing means for securing the top cover,

Figure 5 is an enlarged detail of said securing means,

Figure 6 is a side elevation, showing the exhaust passage clip, and

Figure 7 is a pictorial view of the clip.

In the embodiment shown in the drawings, a cylinder 1, open at both ends, has two pistons 2 slidably mounted therein. Each piston has a dat face at either end separated by the length of the piston and the outer faces 3 of the pistons coact with the outer ends -b of the cylinder to form a scavenging or charging air pump for the engine. Since there is no restriction en the diameter of the outer faces 3 of the pistons and the outer ends 1b of the cylinder imposed by the construction of the engine, an improved performance of the engine is possible if the outer faces 3 of the pistons are made somewhat larger in diameter than the working face 4 of the pistons 2, as shown in Figure l of the drawings. The relative proportions of the outer faces 3 and the working faces 4 of the pistons are of course a matter for the designer of a particular engine to give a desired scavenging or charging air pressure.

The part la of the cylinder and the working faces 4 f the pistons are designed relative to each other to have a common working clearance. A plurality of piston rings 2a are provided in circumferential grooves near the working face i of the piston, in the well known way for the usual purpose. rEhe surface of the piston 2b adjacent to the working face 4 of the piston ccacts with the cylinder wall to form a front guide means to assist in guiding each piston while it reciprocates. The wcrl.- ing faces i of the pistons and the part of the cylinder 1a surrounding the working faces 4 and between said working faces form the combustion chamber of the engine.

Further piston rings 3a are provided in circumfertial grooves near the outer faces 3 of the pistons in the well known way. These piston rings 3a assist in increasing the efficiency of the air pump and the surface of the piston 3b adjacent to the piston rings 3a coacts with the cylinder wall to form a rear guide means to assist the front guide means in guiding each piston as it reciprocates, and for this latter purpose this surface 3b is spaced an appropriate distance from the working face 4 of the piston, for example the distance shown particularly in Figure l.

In the kinematic chain connecting each piston to the driven member or members, the latter comprising a rotating crank shaft 19, there is a gudgeon pin S pivot-ed in a bush 6 in each piston 2, the pivotal axis being transverse to the longitudinal axis of the piston and intermediate of the length of the piston, being substantially equidistant from the piston rings 2a and 3a, and being on a diameter of the piston. The gudgeon pin 5 is made of such a size as to allow a cylindrical aperture 7 to be bored therein (as shown particularly in Figure 3).

An oscillating lever 9 is associated with the piston Z by having a cylindrical end 3 having a bronze bush 10 which is slidably ted within the aperture 7 of the gudgeon pin, the aperture 7 being normal to the longitudinal axis of the gudgeon pin 5.

The bronze bush 10 has a spiral oil groove (not shown) on its outer curved surface, which oil groove is fed from the oil bore 11 and the space 13. The bush 6 also has a spiral oil groove (not shown) cut in its curved surface which communicates with both the space 13 and an oil 3 bore 12 leading to an orifice 14, from which oil is squirted against the rear face 15 of the piston 2, thus aiding its cooling. The oil then drains to the crankcase 16a of the engine frame 16 (which crankcase is adapted to receive the oil) through a slot 17 in the wall of each piston and through slots 18 in the cylinder wall.

Each oscillating lever 9 also extends from the gudgeon pin 5 through the slot 17 in the piston and the slot 18 in the cylinder wall, into the crankcase 16a. The oscillating lever has of course a working clearance from the edges of the slots 17 and 18.

The remaining links in each kinematic chain connecting a piston to the rotating crankshaft 19, comprise a connecting rod 20 and a crank 21, the latter forming part of the crankshaft 19. The connecting rods 20 and cranks 21 are shown only diagrammatically in Figure 1. Each oscillating lever 9 fulcrums on a short shaft 22 on which it oscillates and the outer ends of the shafts 22 are mounted in stress members 23 which are arranged one on either side of the oscillating lever. The stress members 23 are of a high tensile steel or other suitable material cast into the engine frame 16 which is of a light alloy, e. g. aluminium.

The lower end 24 of each rocker arm has fitted thereto a stub shaft 25 which is parallel to the short shaft 22 and the connecting rod 20 movably connects each oscillating lever 9 to a crank 21 of the crankshaft 19, a pair of substantially opposed cranks 21 being provided.

The cranks may be at a slightly different angle from 180 to each other to enable the piston which uncovers the exhaust ports 40 to do this before the inlet ports 43 are uncovered. Other means of achieving this may of course be provided, for example, by lowering the axis of the crankshaft 19 below the line joining the axes of the pins 25 when the pistons are in their mid position.

The crankshaft 19 rotates in bearings 26-26a mounted in the engine frame 16, an oil seal 27 being provided as shown. Oil is contained in the lower part of the crankcase in the well known way as mentioned above, and is pumped therefrom by an oil pump (not shown) through a series of channels 84 and 20a in the crankshaft 19, and the connecting rods 20, the oil bore 11 in the oscillating levers 9, the space 13 and oil hole 12 in the pistons 2. These oil channels connect through orifices in the various links of the kinematic chain with the coacting surfaces interconnecting said links, the bearings themselves having suitable oil grooves cut in their surfaces. Some of the oil lubricates the bearings and some passes on to be squirted from the orifice 14 as described above.

A flywheel 28 is provided, driven by the crankshaft 19 and of suicient inertia to enable its momentum when rotated at a sutiicient speed, to return the pistons to the position where they receive the next power impulse. 4The flywheel also of course lessens the cyclical speed variations in the known manner.

The two oscillating levers 9 are allochirally similar, being so arranged that if the stub shaft 25 of the right hand rocker arm is on the same side as the viewer in Figure l, the stub shaft 25 of the left hand oscillating lever is on the opposite side to the viewer. Because the cranks 21 are opposed, when the crankshaft rotates, the two opposed pistons in each cylinder reciprocate, being guided in the cylinder by the coacting of the piston surfaces 2b and 3b with the cylinder wall, the pistons alternately approaching to and receding from each other.

The flywheel 28 has radial vanes 29 fitted thereto and the housing 30 for the flywheel (integral with the crankcase) has air outlets 31 in its periphery. An end cover 30a on the housing 30 has a V-shaped ange 32 and the housing 30 has a corresponding V-shaped flange 33. The end cover 30a is held in place by a clamp ring 34 whichV has suitable quick release means (not shown) for holding it in operative position on the anges 32 and 33.'

The -vanes 2 9 and the housing 30 form a cooling air Y 4 pump for circulating cooling air through the engine. The cooling air pump draws air through a finned passage 36 leading from the outer air through the crankcase 16u. The fins 37 and 37a fitted both within the crankcase and within the passage 36, assist in transferring heat from the lubricating oil in the crankcase to the air passing through the passage 36 to assist in cooling the engine. The air pump also draws air through an inlet 35 communicating with a chamber 38 surrounding the central portion of the cylinder 1 which is nned as at 39.

The chamber 38 communicates with the outside air through suitably placed apertures, and baffles (not shown) are provided to direct airflow over the top of the cylinder.

The cylinder is provided with a series of apertures 40 (already mentioned) which form the ports for the exhaust gases, the exhaust ports communicating with an exhaust ring 41 leading to an exhaust pipe 42. The exhaust ring is in two portions held by suitable clamps 41a (Figures 6 and 7) engaging lugs 41b on the exhaust ring 41 and lugs 42a on the exhaust pipe 42 to enable it to be removed and replaced in its position.

A further series of apertures 43 form the inlet ports for the scavenging and charging air which communicate with a circumferential groove 44 (Figure l), the side walls 45 of which are integral with the cylinder 1. A short cylinder 46 tits over the walls 45 and an aperture or apertures 47 are cut in a part only of the short cylinder 46 as will be referred to later.

A fuel injection equipment 48 which may be of a suitable known type is provided.

The engine frame 16 is adapted to support and locate the cylinder, as will now be described.

The engine frame 16 extends upwardly to the plane 49 only, this plane being a horizontal plane passing through the longitudinal axis and on a diameter of the cylinder. To support the cylinder in the upper part of the engine frame 16, ribs 50, 51, 52 and 53 are provided which coact with the cylinder as will now be described. The ribs 52 and 53 are each machined to provide a semi-circular depression of the same diameter as the outer diameter of the outer ends 54 of the cylinder. A groove is machined in both the outer ends of the cylinder and in the ribs 52 and 53 to take a rubber or the like resilient sealing ring 55. The rib 50 is machined to provide a semi-circular depression of the same diameter as the outer diameter of the short cylinder 46, the rib 50 and the cylinder 46 being each grooved to accommodate a sealing ring 56. The rib 51 is machined to provide a semi-circular depression the same diameter as the rib 50 and a flange 57 on the cylinder 1 is machined to this same diameter, the ange 57 and rib 51 being grooved to accommodate a sealing ring 56.

The centres of the semi-circular depressions of course lie on a line which coincides with the longitudinal axial centre line of the cylinder 1 when the latter is in place in the semi-circular depressions.

To locate Vthe cylinder longitudinally in the engine frame 16 a spigot 58 lits into holes drilled in both the cylinder wall and the engine frame 16, as shown. This allows thermal expansion and contraction of the cylinder 1 longitudinally, whilst preventing its movement as a whole. The spigot 58 is substantially the same axial distance longitudinally from either end of the cylinder 1. The spigot 58 may of course be fixed to either the engine frame 16 or to the cylinder 1. Other means to locate the cylinder may be provided, for example, either the cylinder, or the engine frame may have a circumferential flange tted thereon to coact with a corresponding groove in the other member.

It is to be noted also that the short shafts 22 are each substantially the same distance from the spigot 58 which is disposed substantially centrally laterally and longitudinally in the engine frame 16. Also as will be seen more particularly in Figure 3, the longitudinal axes of the oscillating levers 9 are on a line which lies on the vertical diameter of the cylinder 1, the rocker arms being therefore symmetrically located in the engine.

To hold the cylinder in position and to fulfil other purposes which will be described later, a top cover 59 is removably attached to the engine frame 16. The top cover has a at lower surface which ts on the plane 49 of the engine frame 16, some form of sealing means being provided between the two contiguous surfaces, for example, paper impregnated with a jointing compound. Four semi-circular sealing surfaces 60, 61, 62 and 63 are provided spaced apart at spacing such as to correspond with the ribs 50, 51, 52 and 53. Each of said surfaces and the cylinder 1 is grooved to take the sealing rings 55 and 56 as in the case of the ribs 50, 51, 52 and 53. The air pump formed by the coacting of the outer faces 3 of the pistons and the outer ends 54 of the cylinder 1 is operatively connected to the atmosphere and the combustion chamber as follows. A passageway 64 for the scavenging and charging air leads from one outer end 54 of the cylinder and a further passageway 65 leads from the other end 54 of the cylinder, both passageways communieating with the groove 44 through the apertures 47. As stated above the apertures 47 are in a'portion only of the periphery of the short cylinder 46 and this portion is within the passageway 64. The outer wall 66 of the top cover 59 forms one wall of both the passageways 64 and 65 and inner walls 67 and 68 complete the walls of the passageways.

An air inlet 69 having a spring loaded disc valve 70 is provided, the valve 70 allowing air to be drawn into the passageways 64 and 65 from the atmosphere but not permitting egress of air from said passageways to the atmosphere. It is to be noted that the seals 55 and 56 prevent the interchange of air between the passageways 64 and 65 and the outer atmosphere, the interior of the crankcase 16a and the cooling space 38.

Referring to Figures 4 and 5, the top cover 59 is removably attached to the engine frame 16, thus holding the cylinder 1 in position, by four linkages or toggle catches, one towards each end of the inside of the top cover on either side thereof. One such catch is shown in Figure 4. Each catch comprises a curved and hooked link 71 pivotally connected by a pivot 71a to a toggle lever 72, pivoted in turn at 72a to a curved link 72b pivoted in turn, at 72C to a link 72d pivotally mounted at 72e to the top cover 59. The hook of the hooked link 71 engages a pin 73 on the engine frame 16. The hooked links 71 are accessible through apertures 74 on each end of the crankcase 16a, the apertures '74 being covered by cover plates 75 (Figure l).

Other means of securing the top cover in position may of course be provided, for example the top cover may be bolted in place.

The cylinder 1 is releasably held in the top cover 59 by suitable means, so that the top cover and the cylinder are removed as a whole, with the pistons still inside the cylinder and with the exhaust ring 41 and exhaust pipe 42 still attached to the cylinder. If it is desired to remove the exhaust ring 41 this may be done at alater stage but because the pistons may be removed from the cylinder once the top cover 59 has been removed from the crankcase 16, removal of the exhaust ring 41 may not be necessary unless it requires replacing.-

So that the piston 2 may be removed endwise from the cylinder 1, flat circular discs 76 (Figure l) are held in grooves 77 in semi-circular apertures in the engine frame 16 and grooves 78 in semi-circular apertures in the top cover 59, seals comprising rubber rings 79 sealing the discs in the said grooves.

The semi-circular apertures in the top cover must have a greater diameter than the largest diameter of the pistons 2 to enable the latter to be withdrawn endwise from the cylinder 1 while the cylinder is still held within the top cover 59.

The fuel injection pump 48 is operated through a rocker arm 80, push rod 81 and cam follower 82 by a cam 83 rotating with the crank shaft 19.

The crankshaft is accessible for withdrawal by removing the caps of the bearings 26, 26a and the members 35 andv 36.

In operation, the crankshaft revolves, the oscillating levers fulcrum on the short shafts 22, thus oscillating over a short arc, and the ends 8 of the oscillating levers 9 slide slightly in the gudgeon pins 5 as the pistons 2 reciprocate, since the ends of the oscillating levers move in an arc of a circle while the gudgeon pins 5 are constrained by the movement of the piston, guided by the piston surfaces 2b and 3b in the cylinder 1, to move in a straight line, the gudgeon pins turning slightly in the bushes 6 in the pistons 2.

It is to be noted that since each link in the kinematic chain is constrained to move over a definite path, the kinematic chain is completely restrained.

Because the end 8 slides in the aperture 7 in the gudgeon pin 5, oil fills the space 13 then when the end 8 moves upwardly the oil is squirted through the oriiice 14 in spurts.

The engine may be dismantled for decarbonizing and the like by releasing the top cover 59 and limiting the said top cover and the cylinder l containing the pistons 2 vertically i. e. along a line at right angles to the longitudinal axis of the cylinder. The ends 8 of the oscillating levers 9 simply slide out of the apertures 7 in the gudgeon pins thus disengaging the pistons from the oscillating levers. The cylinder may then be moved clear of the engine frame and the pistons removed endwise from the cylinder, which may be left attached to the top cover.

Because of the sliding of the ends 8 of the oscillating levers in the gudgeon pins 5 and because of the rotation of the gudgeon pins in the piston bosses, the distance of each end of the cylinder above the crankshaft is not critical and some misalignment of the cylinder relative to the crankshaft in this respect may be accommodated without straining the engine.

The stress members 23 take the main dynamic stresses substantially linearly when the engine is running.

Suitable cast materials may be used throughout in the manufacture of the engine or the engine may be built up by welding or the like from rolled or other suitable mateh rials, or a mixture of the two may be used.

In the preferred form, the construction is such that the exhaust ports 4@ are nearer the longitudinal centre (on which the pin 53 lies) than the inlet ports 43, so that the exhaust ports maybe uncovered first on outward strokes of the pistons. With this construction, the direction of rotation of the engine may be reversed if desired, by reversing the cylinder 1 and top cover 59 end for end in the engine frame 16 and adjusting the cam 83 operating the fuel injector in order to inject the fuel at the correct time.

The axial centre of the gudgeon pin need not necessarily lie on a diameter of the piston, but may be slightly below it in order to enable the 'rocker arm to be shortened slightly;

It is of course obvious that although throughout this specification of which the claims .form part, the engine shown has been described as though the cylinder is in a horizontal plane with the crankshaft below it, other juxtapositions of these two members are possible, for example the cylinder may be vertically disposed or may be disposed horizontally below the crankshaft.

Although a single cylinder engine has been described in detail, it is obvious that the invention may be applied to a multi-cylinder engine.

Although scavenging and charging air has been provided in the above-described engine by the coacting of the outer faces 3 of the pistons 2 with the outer ends of the cylinder 1, any engineer skilled in the art would be able to adapt the engine as described to use other known means to provide such air. For example, a rotary blower may be used, in which case the cylinder 1 would be of the same internal diameter throughout its length and the piston rings 3a would be replaced by a single oil ring near the outer end of the skirt of the piston.

Also, although air cooling of the cylinder has been described, an engineer` skilled in the art could readily adapt the engine to water cooling.

We claim:

l. In an internal combustion engine of the type having a. cylinder, a pair of double acting opposed pistons in said cylinder, a gudgeon pin in each said piston, a connecting rod lever for each said piston slidably insertable into said gudgeon pin, a frame member to support said connecting rod levers and said cylinder, a cover removably attached to said frame member, said cover having a at lower surface for engagement with the said frame member to hold the said cylinder against motion, a plurality of sealing rings disposed between respective points of engagement, and a plurality of toggle catches to removably secure the said cover to the said frame whereby upon release of said toggle catches the said cylinder may be dismounted, said piston thereby removing from slidable engagement with said connecting rod.

2. The invention as set forth in claim l, wherein the respective said toggle catches comprise a rst hooked link, a toggle lever pivotally connected to said first hooked link, a second link pivotally connected to said toggle lever, a third link pivotally connected to said second link, said third link adapted to be pivotally connected to said cover, and a pin member secured to said frame member to be embraced by said first hooked link.

3. In an internal combustion engine of the type having an elongated cylinder and a plurality of pistons slidable therein, a gudgeon pin centrally mounted in each of said pistons, a frame member adapted to support said cylinder, means to locate said cylinder against endwise movement in said frame, kinematic chains in said frame including lever means slidably engaging said gudgeon pins in said pistons to transmit force therefrom and thereto, and lever operated tension means releasably secured to said engine frame to hold said cylinder against vertical motion whereby complete renovation of the engine may be made by release of said clamping means after which said cylinder is removable from said engine frame enabling disconnection of said pistons from said kinematic chains and quick replacement thereof by new parts.

4. In an internal combustion engine of the type having an elongated cylinder and a plurality of double acting pistons slidable therein, a gudgeon pin centrally mounted in each said piston, a frame member to support said cylinder, a lever means for each said piston connected' Within said frame to a connecting rod and crank having one end projecting into said cylinder to operatively engage said gudgeon pin, a cover removably attached to said frame member, said cover having a at lower surface for engagement with the said frame member to hold the said cylinder against vertical motion, a plurality of sealing an elongated cylinder and a-plurality'of double acting pistons slidably received therein, a gudgeon pin mounted centrally in each said piston, a frame member adapted to support said cylinder, lever means for each said piston mounted in said frame and projecting into said cylinder to engage said gudgeon pin, a cover lremovably attached to said frame member, said cover having a at lower surface for engagement with said frame member and said cylinder to hold the same against vertical motion, a plurality of sealing rings disposed between respectivepoints of engagement, and releasable clamping means to removably secure the said cover to the said frame whereby upon removal of said cover the said cylinder may be dismounted without use of wrenches or other tools.

6, In an internal combustion engine of the type'having an elongated cylinder and a plurality of pistons slidable therein, a frame member adapted to support said cylinder, lever means in said frame extending into said cylinder to engage said pistons to transmit forces from and to said pistons, a cover removably attached to said frame member, cylinder air inlet passages formed in said cover, said cover having a flat lower surface'for engagement with said frame member and concave surfaces to engage `said cylinder to hold the same against vertical motion, a plurality of sealing rings disposed in respective planes of engagement, and releasable clamping means to removably secure the said' cover to the said frame to permit rapid dismounting thereof without the use of tools.

' 7.I In an' internal combustion engine comprising a frame, 'a cover, a cylinder open at both ends cradled in arcuately concave webs of said frame and of said cover, two pistons slidably mounted in said cylinder each said 'piston presenting a Working face at each end separated by the length of said piston, the working faces presented inwardly toward each other dening the ends of the combustion chamber of the engine and the working faces presented outwardly coacting with a portion of the said frame and of said cover to form a scavenging air pump, means to drivingly connect said pistons to a rotatable shaft said means including a lever opposite each piston pivoted externally of said cylinder and extending into said cylinder to engage said pistons, and lever means to secure said cover to said frame said lever means being so constructed and arranged that securement or release is made without use Vof tools, whereby removal and replacement of said cylinder and pistons may be quickly performed.

References Cited in the le of this patent UNITED STATES PATENTS

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2026-09-01T19:52:21Z

Toro Andersen:

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2026-09-01T19:51:27Z

Toro Andersen:

Morton patent

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2026-09-01T19:48:02Z

Toro Andersen:

Digby B. Morton patent drawing page 1

Digby B. Morton

2026-09-01T19:42:45Z

Toro Andersen:

Morton patented a opposed piston engine in the 1950s.

He was situated in Red Hill, Papakura, in New Zealand.

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Jan. 10, 1956 D. B. MORTON ET AL INTERNAL-COMBUSTION ENGINES `I5 Sheets-Sheet l

Filed April 20, 1953 Inventors D. 13H optan? 51. llozozz/Jfp Jan. 10, 1956 D. B. MoRToN ET AL INTERNAL-COMBUSTION ENGINES 5 Sheets-Sheet 2 Filed April 20, 1953 Jan. 10, 1956 D. B. MORTON ET AL 2,730,087

INTERNAL-COMBUSTION ENGINES Filed April 20, 1953 3 Sheets-Sheet 3 F/GZ nted States Patent O INTERNAL-COMBUSTION ENGINES Digby B. Morton, Digby B. Morton, Jr., and Graeme E. Morton, Kilbirnie, Wellington, New Zealand Application April 20, 1953, Serial No. 349,834

Claims priority, application New Zealand May Z, 1952 7 Claims. (Cl. 123-51) This invention relates to reciprocating internal combustion engines of the opposed piston type wherein the power is transmitted from each opposed piston through a kinematic chain to a driven member or members, such, for example, as a common rotating crank shaft; and is particularly though not solely of use in two cycle compression ignition engines of this type.

It is known that internal combustion engines of the type described possess thermodynamic advantages, particularly in the case of two cycle compression ignition engines of this type, but in present engines of the type described, disadvantages exist in that the kinematic chains connecting the pistons to the driven member or members have disadvantages, for example being unduly complicated or expensive to manufacture. Furthermore, such engines are not compact and must be built larger and heavier than would otherwise be necessary in order to accommodate the kinematic chains.

lt is therefore an object of the present invention to provide an internal combustion engine and preferably a two cycle compression ignition engine, of the type described, which is simple in construction and therefore cheap to manufacture, and which is compact in size compared with present engines of this type.

It is a further object of the present invention to provide an internal combustion engine of the type described which may be more easily dismantled and reassembled than present engines of this type.

lt is a still further object of the present invention in a preferred form to provide an internal combustion engine in which scavenging or charging air may be provided in sufficient quantity to give eicient scavenging and charging in a simple and convenient manner.

Accordingly in one aspect the invention consists in an internal combustion engine of the type described, wherein the kinematic chain connecting each piston to the driven member or members includes an oscillating lever which is so slidably associated with said piston that said lever will oscillate when said piston reciprocates.

In a further aspect the invention consists in an internal combustion engine of the type described, wherein the kinematic chain connecting each piston to the driven member or members includes an oscillating lever, which is so associated with said piston intermediate of the length of the piston, that said lever will oscillate when said piston reciprocates.

An advantage of the invention in its preferred form resides in the automatic allowance for misalignment between the cylinder and the crankshaft by the means used to connect the pistons, the gudgeon pins and the rocker arms. Y

One preferred form of the invention will now be described with reference to the accompanying drawings in which:

Figure l is a vertical cross-section on the line 1 1, Figure 2, of an engine constructed according to the invention, Y n

Figure 2 is a vertical cross-section on the line 2-2,

Figure l, the connecting rods being omitted and some parts being shown in elevation,

Figure 13 is a section of a piston and its associated oscillating lever on the lines 3 3, Figure l,

Figure 4 is a fragmentary side view, partly in broken lines, showing the securing means for securing the top cover,

Figure 5 is an enlarged detail of said securing means,

Figure 6 is a side elevation, showing the exhaust passage clip, and

Figure 7 is a pictorial view of the clip.

In the embodiment shown in the drawings, a cylinder 1, open at both ends, has two pistons 2 slidably mounted therein. Each piston has a dat face at either end separated by the length of the piston and the outer faces 3 of the pistons coact with the outer ends -b of the cylinder to form a scavenging or charging air pump for the engine. Since there is no restriction en the diameter of the outer faces 3 of the pistons and the outer ends 1b of the cylinder imposed by the construction of the engine, an improved performance of the engine is possible if the outer faces 3 of the pistons are made somewhat larger in diameter than the working face 4 of the pistons 2, as shown in Figure l of the drawings. The relative proportions of the outer faces 3 and the working faces 4 of the pistons are of course a matter for the designer of a particular engine to give a desired scavenging or charging air pressure.

The part la of the cylinder and the working faces 4 f the pistons are designed relative to each other to have a common working clearance. A plurality of piston rings 2a are provided in circumferential grooves near the working face i of the piston, in the well known way for the usual purpose. rEhe surface of the piston 2b adjacent to the working face 4 of the piston ccacts with the cylinder wall to form a front guide means to assist in guiding each piston while it reciprocates. The wcrl.- ing faces i of the pistons and the part of the cylinder 1a surrounding the working faces 4 and between said working faces form the combustion chamber of the engine.

Further piston rings 3a are provided in circumfertial grooves near the outer faces 3 of the pistons in the well known way. These piston rings 3a assist in increasing the efficiency of the air pump and the surface of the piston 3b adjacent to the piston rings 3a coacts with the cylinder wall to form a rear guide means to assist the front guide means in guiding each piston as it reciprocates, and for this latter purpose this surface 3b is spaced an appropriate distance from the working face 4 of the piston, for example the distance shown particularly in Figure l.

In the kinematic chain connecting each piston to the driven member or members, the latter comprising a rotating crank shaft 19, there is a gudgeon pin S pivot-ed in a bush 6 in each piston 2, the pivotal axis being transverse to the longitudinal axis of the piston and intermediate of the length of the piston, being substantially equidistant from the piston rings 2a and 3a, and being on a diameter of the piston. The gudgeon pin 5 is made of such a size as to allow a cylindrical aperture 7 to be bored therein (as shown particularly in Figure 3).

An oscillating lever 9 is associated with the piston Z by having a cylindrical end 3 having a bronze bush 10 which is slidably ted within the aperture 7 of the gudgeon pin, the aperture 7 being normal to the longitudinal axis of the gudgeon pin 5.

The bronze bush 10 has a spiral oil groove (not shown) on its outer curved surface, which oil groove is fed from the oil bore 11 and the space 13. The bush 6 also has a spiral oil groove (not shown) cut in its curved surface which communicates with both the space 13 and an oil 3 bore 12 leading to an orifice 14, from which oil is squirted against the rear face 15 of the piston 2, thus aiding its cooling. The oil then drains to the crankcase 16a of the engine frame 16 (which crankcase is adapted to receive the oil) through a slot 17 in the wall of each piston and through slots 18 in the cylinder wall.

Each oscillating lever 9 also extends from the gudgeon pin 5 through the slot 17 in the piston and the slot 18 in the cylinder wall, into the crankcase 16a. The oscillating lever has of course a working clearance from the edges of the slots 17 and 18.

The remaining links in each kinematic chain connecting a piston to the rotating crankshaft 19, comprise a connecting rod 20 and a crank 21, the latter forming part of the crankshaft 19. The connecting rods 20 and cranks 21 are shown only diagrammatically in Figure 1. Each oscillating lever 9 fulcrums on a short shaft 22 on which it oscillates and the outer ends of the shafts 22 are mounted in stress members 23 which are arranged one on either side of the oscillating lever. The stress members 23 are of a high tensile steel or other suitable material cast into the engine frame 16 which is of a light alloy, e. g. aluminium.

The lower end 24 of each rocker arm has fitted thereto a stub shaft 25 which is parallel to the short shaft 22 and the connecting rod 20 movably connects each oscillating lever 9 to a crank 21 of the crankshaft 19, a pair of substantially opposed cranks 21 being provided.

The cranks may be at a slightly different angle from 180 to each other to enable the piston which uncovers the exhaust ports 40 to do this before the inlet ports 43 are uncovered. Other means of achieving this may of course be provided, for example, by lowering the axis of the crankshaft 19 below the line joining the axes of the pins 25 when the pistons are in their mid position.

The crankshaft 19 rotates in bearings 26-26a mounted in the engine frame 16, an oil seal 27 being provided as shown. Oil is contained in the lower part of the crankcase in the well known way as mentioned above, and is pumped therefrom by an oil pump (not shown) through a series of channels 84 and 20a in the crankshaft 19, and the connecting rods 20, the oil bore 11 in the oscillating levers 9, the space 13 and oil hole 12 in the pistons 2. These oil channels connect through orifices in the various links of the kinematic chain with the coacting surfaces interconnecting said links, the bearings themselves having suitable oil grooves cut in their surfaces. Some of the oil lubricates the bearings and some passes on to be squirted from the orifice 14 as described above.

A flywheel 28 is provided, driven by the crankshaft 19 and of suicient inertia to enable its momentum when rotated at a sutiicient speed, to return the pistons to the position where they receive the next power impulse. 4The flywheel also of course lessens the cyclical speed variations in the known manner.

The two oscillating levers 9 are allochirally similar, being so arranged that if the stub shaft 25 of the right hand rocker arm is on the same side as the viewer in Figure l, the stub shaft 25 of the left hand oscillating lever is on the opposite side to the viewer. Because the cranks 21 are opposed, when the crankshaft rotates, the two opposed pistons in each cylinder reciprocate, being guided in the cylinder by the coacting of the piston surfaces 2b and 3b with the cylinder wall, the pistons alternately approaching to and receding from each other.

The flywheel 28 has radial vanes 29 fitted thereto and the housing 30 for the flywheel (integral with the crankcase) has air outlets 31 in its periphery. An end cover 30a on the housing 30 has a V-shaped ange 32 and the housing 30 has a corresponding V-shaped flange 33. The end cover 30a is held in place by a clamp ring 34 whichV has suitable quick release means (not shown) for holding it in operative position on the anges 32 and 33.'

The -vanes 2 9 and the housing 30 form a cooling air Y 4 pump for circulating cooling air through the engine. The cooling air pump draws air through a finned passage 36 leading from the outer air through the crankcase 16u. The fins 37 and 37a fitted both within the crankcase and within the passage 36, assist in transferring heat from the lubricating oil in the crankcase to the air passing through the passage 36 to assist in cooling the engine. The air pump also draws air through an inlet 35 communicating with a chamber 38 surrounding the central portion of the cylinder 1 which is nned as at 39.

The chamber 38 communicates with the outside air through suitably placed apertures, and baffles (not shown) are provided to direct airflow over the top of the cylinder.

The cylinder is provided with a series of apertures 40 (already mentioned) which form the ports for the exhaust gases, the exhaust ports communicating with an exhaust ring 41 leading to an exhaust pipe 42. The exhaust ring is in two portions held by suitable clamps 41a (Figures 6 and 7) engaging lugs 41b on the exhaust ring 41 and lugs 42a on the exhaust pipe 42 to enable it to be removed and replaced in its position.

A further series of apertures 43 form the inlet ports for the scavenging and charging air which communicate with a circumferential groove 44 (Figure l), the side walls 45 of which are integral with the cylinder 1. A short cylinder 46 tits over the walls 45 and an aperture or apertures 47 are cut in a part only of the short cylinder 46 as will be referred to later.

A fuel injection equipment 48 which may be of a suitable known type is provided.

The engine frame 16 is adapted to support and locate the cylinder, as will now be described.

The engine frame 16 extends upwardly to the plane 49 only, this plane being a horizontal plane passing through the longitudinal axis and on a diameter of the cylinder. To support the cylinder in the upper part of the engine frame 16, ribs 50, 51, 52 and 53 are provided which coact with the cylinder as will now be described. The ribs 52 and 53 are each machined to provide a semi-circular depression of the same diameter as the outer diameter of the outer ends 54 of the cylinder. A groove is machined in both the outer ends of the cylinder and in the ribs 52 and 53 to take a rubber or the like resilient sealing ring 55. The rib 50 is machined to provide a semi-circular depression of the same diameter as the outer diameter of the short cylinder 46, the rib 50 and the cylinder 46 being each grooved to accommodate a sealing ring 56. The rib 51 is machined to provide a semi-circular depression the same diameter as the rib 50 and a flange 57 on the cylinder 1 is machined to this same diameter, the ange 57 and rib 51 being grooved to accommodate a sealing ring 56.

The centres of the semi-circular depressions of course lie on a line which coincides with the longitudinal axial centre line of the cylinder 1 when the latter is in place in the semi-circular depressions.

To locate Vthe cylinder longitudinally in the engine frame 16 a spigot 58 lits into holes drilled in both the cylinder wall and the engine frame 16, as shown. This allows thermal expansion and contraction of the cylinder 1 longitudinally, whilst preventing its movement as a whole. The spigot 58 is substantially the same axial distance longitudinally from either end of the cylinder 1. The spigot 58 may of course be fixed to either the engine frame 16 or to the cylinder 1. Other means to locate the cylinder may be provided, for example, either the cylinder, or the engine frame may have a circumferential flange tted thereon to coact with a corresponding groove in the other member.

It is to be noted also that the short shafts 22 are each substantially the same distance from the spigot 58 which is disposed substantially centrally laterally and longitudinally in the engine frame 16. Also as will be seen more particularly in Figure 3, the longitudinal axes of the oscillating levers 9 are on a line which lies on the vertical diameter of the cylinder 1, the rocker arms being therefore symmetrically located in the engine.

To hold the cylinder in position and to fulfil other purposes which will be described later, a top cover 59 is removably attached to the engine frame 16. The top cover has a at lower surface which ts on the plane 49 of the engine frame 16, some form of sealing means being provided between the two contiguous surfaces, for example, paper impregnated with a jointing compound. Four semi-circular sealing surfaces 60, 61, 62 and 63 are provided spaced apart at spacing such as to correspond with the ribs 50, 51, 52 and 53. Each of said surfaces and the cylinder 1 is grooved to take the sealing rings 55 and 56 as in the case of the ribs 50, 51, 52 and 53. The air pump formed by the coacting of the outer faces 3 of the pistons and the outer ends 54 of the cylinder 1 is operatively connected to the atmosphere and the combustion chamber as follows. A passageway 64 for the scavenging and charging air leads from one outer end 54 of the cylinder and a further passageway 65 leads from the other end 54 of the cylinder, both passageways communieating with the groove 44 through the apertures 47. As stated above the apertures 47 are in a'portion only of the periphery of the short cylinder 46 and this portion is within the passageway 64. The outer wall 66 of the top cover 59 forms one wall of both the passageways 64 and 65 and inner walls 67 and 68 complete the walls of the passageways.

An air inlet 69 having a spring loaded disc valve 70 is provided, the valve 70 allowing air to be drawn into the passageways 64 and 65 from the atmosphere but not permitting egress of air from said passageways to the atmosphere. It is to be noted that the seals 55 and 56 prevent the interchange of air between the passageways 64 and 65 and the outer atmosphere, the interior of the crankcase 16a and the cooling space 38.

Referring to Figures 4 and 5, the top cover 59 is removably attached to the engine frame 16, thus holding the cylinder 1 in position, by four linkages or toggle catches, one towards each end of the inside of the top cover on either side thereof. One such catch is shown in Figure 4. Each catch comprises a curved and hooked link 71 pivotally connected by a pivot 71a to a toggle lever 72, pivoted in turn at 72a to a curved link 72b pivoted in turn, at 72C to a link 72d pivotally mounted at 72e to the top cover 59. The hook of the hooked link 71 engages a pin 73 on the engine frame 16. The hooked links 71 are accessible through apertures 74 on each end of the crankcase 16a, the apertures '74 being covered by cover plates 75 (Figure l).

Other means of securing the top cover in position may of course be provided, for example the top cover may be bolted in place.

The cylinder 1 is releasably held in the top cover 59 by suitable means, so that the top cover and the cylinder are removed as a whole, with the pistons still inside the cylinder and with the exhaust ring 41 and exhaust pipe 42 still attached to the cylinder. If it is desired to remove the exhaust ring 41 this may be done at alater stage but because the pistons may be removed from the cylinder once the top cover 59 has been removed from the crankcase 16, removal of the exhaust ring 41 may not be necessary unless it requires replacing.-

So that the piston 2 may be removed endwise from the cylinder 1, flat circular discs 76 (Figure l) are held in grooves 77 in semi-circular apertures in the engine frame 16 and grooves 78 in semi-circular apertures in the top cover 59, seals comprising rubber rings 79 sealing the discs in the said grooves.

The semi-circular apertures in the top cover must have a greater diameter than the largest diameter of the pistons 2 to enable the latter to be withdrawn endwise from the cylinder 1 while the cylinder is still held within the top cover 59.

The fuel injection pump 48 is operated through a rocker arm 80, push rod 81 and cam follower 82 by a cam 83 rotating with the crank shaft 19.

The crankshaft is accessible for withdrawal by removing the caps of the bearings 26, 26a and the members 35 andv 36.

In operation, the crankshaft revolves, the oscillating levers fulcrum on the short shafts 22, thus oscillating over a short arc, and the ends 8 of the oscillating levers 9 slide slightly in the gudgeon pins 5 as the pistons 2 reciprocate, since the ends of the oscillating levers move in an arc of a circle while the gudgeon pins 5 are constrained by the movement of the piston, guided by the piston surfaces 2b and 3b in the cylinder 1, to move in a straight line, the gudgeon pins turning slightly in the bushes 6 in the pistons 2.

It is to be noted that since each link in the kinematic chain is constrained to move over a definite path, the kinematic chain is completely restrained.

Because the end 8 slides in the aperture 7 in the gudgeon pin 5, oil fills the space 13 then when the end 8 moves upwardly the oil is squirted through the oriiice 14 in spurts.

The engine may be dismantled for decarbonizing and the like by releasing the top cover 59 and limiting the said top cover and the cylinder l containing the pistons 2 vertically i. e. along a line at right angles to the longitudinal axis of the cylinder. The ends 8 of the oscillating levers 9 simply slide out of the apertures 7 in the gudgeon pins thus disengaging the pistons from the oscillating levers. The cylinder may then be moved clear of the engine frame and the pistons removed endwise from the cylinder, which may be left attached to the top cover.

Because of the sliding of the ends 8 of the oscillating levers in the gudgeon pins 5 and because of the rotation of the gudgeon pins in the piston bosses, the distance of each end of the cylinder above the crankshaft is not critical and some misalignment of the cylinder relative to the crankshaft in this respect may be accommodated without straining the engine.

The stress members 23 take the main dynamic stresses substantially linearly when the engine is running.

Suitable cast materials may be used throughout in the manufacture of the engine or the engine may be built up by welding or the like from rolled or other suitable mateh rials, or a mixture of the two may be used.

In the preferred form, the construction is such that the exhaust ports 4@ are nearer the longitudinal centre (on which the pin 53 lies) than the inlet ports 43, so that the exhaust ports maybe uncovered first on outward strokes of the pistons. With this construction, the direction of rotation of the engine may be reversed if desired, by reversing the cylinder 1 and top cover 59 end for end in the engine frame 16 and adjusting the cam 83 operating the fuel injector in order to inject the fuel at the correct time.

The axial centre of the gudgeon pin need not necessarily lie on a diameter of the piston, but may be slightly below it in order to enable the 'rocker arm to be shortened slightly;

It is of course obvious that although throughout this specification of which the claims .form part, the engine shown has been described as though the cylinder is in a horizontal plane with the crankshaft below it, other juxtapositions of these two members are possible, for example the cylinder may be vertically disposed or may be disposed horizontally below the crankshaft.

Although a single cylinder engine has been described in detail, it is obvious that the invention may be applied to a multi-cylinder engine.

Although scavenging and charging air has been provided in the above-described engine by the coacting of the outer faces 3 of the pistons 2 with the outer ends of the cylinder 1, any engineer skilled in the art would be able to adapt the engine as described to use other known means to provide such air. For example, a rotary blower may be used, in which case the cylinder 1 would be of the same internal diameter throughout its length and the piston rings 3a would be replaced by a single oil ring near the outer end of the skirt of the piston.

Also, although air cooling of the cylinder has been described, an engineer` skilled in the art could readily adapt the engine to water cooling.

We claim:

l. In an internal combustion engine of the type having a. cylinder, a pair of double acting opposed pistons in said cylinder, a gudgeon pin in each said piston, a connecting rod lever for each said piston slidably insertable into said gudgeon pin, a frame member to support said connecting rod levers and said cylinder, a cover removably attached to said frame member, said cover having a at lower surface for engagement with the said frame member to hold the said cylinder against motion, a plurality of sealing rings disposed between respective points of engagement, and a plurality of toggle catches to removably secure the said cover to the said frame whereby upon release of said toggle catches the said cylinder may be dismounted, said piston thereby removing from slidable engagement with said connecting rod.

2. The invention as set forth in claim l, wherein the respective said toggle catches comprise a rst hooked link, a toggle lever pivotally connected to said first hooked link, a second link pivotally connected to said toggle lever, a third link pivotally connected to said second link, said third link adapted to be pivotally connected to said cover, and a pin member secured to said frame member to be embraced by said first hooked link.

3. In an internal combustion engine of the type having an elongated cylinder and a plurality of pistons slidable therein, a gudgeon pin centrally mounted in each of said pistons, a frame member adapted to support said cylinder, means to locate said cylinder against endwise movement in said frame, kinematic chains in said frame including lever means slidably engaging said gudgeon pins in said pistons to transmit force therefrom and thereto, and lever operated tension means releasably secured to said engine frame to hold said cylinder against vertical motion whereby complete renovation of the engine may be made by release of said clamping means after which said cylinder is removable from said engine frame enabling disconnection of said pistons from said kinematic chains and quick replacement thereof by new parts.

4. In an internal combustion engine of the type having an elongated cylinder and a plurality of double acting pistons slidable therein, a gudgeon pin centrally mounted in each said piston, a frame member to support said cylinder, a lever means for each said piston connected' Within said frame to a connecting rod and crank having one end projecting into said cylinder to operatively engage said gudgeon pin, a cover removably attached to said frame member, said cover having a at lower surface for engagement with the said frame member to hold the said cylinder against vertical motion, a plurality of sealing an elongated cylinder and a-plurality'of double acting pistons slidably received therein, a gudgeon pin mounted centrally in each said piston, a frame member adapted to support said cylinder, lever means for each said piston mounted in said frame and projecting into said cylinder to engage said gudgeon pin, a cover lremovably attached to said frame member, said cover having a at lower surface for engagement with said frame member and said cylinder to hold the same against vertical motion, a plurality of sealing rings disposed between respectivepoints of engagement, and releasable clamping means to removably secure the said cover to the said frame whereby upon removal of said cover the said cylinder may be dismounted without use of wrenches or other tools.

6, In an internal combustion engine of the type'having an elongated cylinder and a plurality of pistons slidable therein, a frame member adapted to support said cylinder, lever means in said frame extending into said cylinder to engage said pistons to transmit forces from and to said pistons, a cover removably attached to said frame member, cylinder air inlet passages formed in said cover, said cover having a flat lower surface'for engagement with said frame member and concave surfaces to engage `said cylinder to hold the same against vertical motion, a plurality of sealing rings disposed in respective planes of engagement, and releasable clamping means to removably secure the said' cover to the said frame to permit rapid dismounting thereof without the use of tools.

' 7.I In an' internal combustion engine comprising a frame, 'a cover, a cylinder open at both ends cradled in arcuately concave webs of said frame and of said cover, two pistons slidably mounted in said cylinder each said 'piston presenting a Working face at each end separated by the length of said piston, the working faces presented inwardly toward each other dening the ends of the combustion chamber of the engine and the working faces presented outwardly coacting with a portion of the said frame and of said cover to form a scavenging air pump, means to drivingly connect said pistons to a rotatable shaft said means including a lever opposite each piston pivoted externally of said cylinder and extending into said cylinder to engage said pistons, and lever means to secure said cover to said frame said lever means being so constructed and arranged that securement or release is made without use Vof tools, whereby removal and replacement of said cylinder and pistons may be quickly performed.

References Cited in the le of this patent UNITED STATES PATENTS

Digby B. Morton

2026-08-31T15:04:19Z

Toro Andersen:

Morton patented a opposed piston engine in the 1950s.

He was situated in Red Hill, Papakura, in New Zealand.

[[File:1957 Digby Morton.png|thumb]]

File:1957 Digby Morton.png

2026-08-31T15:04:10Z

Toro Andersen:

Oil Engine and Gas Turbine April 1957

Digby B. Morton

2026-08-31T15:03:31Z

Toro Andersen: Created page with "Morton patented a opposed piston engine in the 1950s. He was situated in Red Hill, Papakura, in New Zealand."

Morton patented a opposed piston engine in the 1950s.

He was situated in Red Hill, Papakura, in New Zealand.

Main Page

2026-07-26T15:54:27Z

Toro Andersen:

__NOTOC__
[[File:Fig_1_Otto_4_stroke_1876.png|right|thumb|Ottos engine]]Welcome to the '''Old Engines Wiki''' - The wiki for old engines, and history in motion, for every kind of engine.

Old engines and the history of internal combustion engines are fascinating topics. We try to do our part in preserving the history of these magnificient machines and their use.

We will try to be the definitive resource for all things related to internal combustion engines! From the earliest prototypes to the most advanced modern designs, this wiki is dedicated to chronicling the fascinating history of these powerful machines that have shaped the modern world.

For over a century, internal combustion engines have been the driving force behind countless innovations in transportation, manufacturing, and power generation. From the first gasoline-powered automobile to the massive diesel engines that power cargo ships, these machines have revolutionized the way we live and work.

But the history of internal combustion engines is more than just a series of technological advancements. It's a story of innovation, competition, and perseverance. It's a tale of brilliant inventors, determined entrepreneurs, and visionary engineers who overcame countless challenges to push the boundaries of what was thought possible.

This wiki is dedicated to preserving this rich legacy for future generations. With in-depth articles, detailed diagrams, and stunning photographs, we aim to provide a comprehensive and engaging resource that celebrates the history of internal combustion engines and the people who made it all possible.

Whether you're a student, an enthusiast, or simply curious about the inner workings of these incredible machines, you'll find a wealth of information and inspiration here. So join us on a journey through time and explore the fascinating history of internal combustion engines!

== We have divided up the oldengin.es website into a few speciality sites. ==
First, this '''WIKI''', that itself is divided into each country and their engine makers. Multi country engine makers are listed in the country where they were founded, or where the headquarters is.

Then we have the [https://forum.oldengin.es forum] that is our social place for discussions about engines, and engine related things. Join us, and learn someting, or even better share your knowledge with fellow engineheads.

We also have a blog, that lives at blog.oldengin.es, where we the websites admins publish out "off topic" things. And a general website at www.oldengin.es that we thing would publish news about the current state of engine building, and stuff that are relevant for the community.

And not least a cloud archive with the content of our library of engine related literature.

We have a large archive of Workshop Manuals, Parts catalogue for some engine manufacturers, mainly German and Swedish. (mtu, Daimler Benz, Maybach Motorenbau, Scania, and Volvo Penta, and a small collection of Detroit Diesel, Twin Disc and some other engine makers. This is available for our supporters eiter as images, or in PDF format.

== Here are some of the tings we have published ==

* [[:Category:Persons|Persons]] - Pages about persons we think are important to our hobby.
*[[:Category:Articles|Articles]] - Pages with Articles from other sources than own research.

== Who are we? ==
[[File:Motorparts.png|left|frameless|254x254px]]
oldengin.es is a domain name, (old engines) and some websites, and is owned by the people who runs [https://mtdb.no Motorteknisk Database] here in Norway. This website is for all the engine makers that dont fit into the mtdb.no platform. We will lean on the mtdb.no lists of engine for the Nordic countries. Also we run the DieselWiki that lives on wiki.cdd.no that mainly cares about Daimler, Benz, Maybach and Zeppelin derived companies and engines.

And it is a way to get some of the history of old engines archived on the internet. We are not doing this to make money, the only thing supporters contributes to is the running costs of this website, and occasional bids on ebay to buy some documentation in high demand.

== Why? ==
About old engines, ask your kid if they can name some old semi-diesel engines? If they google it, we hope we can be of help. The goal is to collect information and make it freely available for the generations that come after us.

== Can i contribute? You need some help? ==
Yes you can, let us make you an account at this engine wiki, and you can publish your knowledge. Mail us at dieselmotor@outlook.com

==Links to interesting places==

https://www.industriegeschichte-ludwigsfelde.de

https://themator.museum-digital.de/ausgabe/showthema.php?m_tid=2056&tid=2056&ver=standalone

File:Motorparts.png

2026-07-26T15:53:58Z

Toro Andersen:

Parts for engine

Studebaker Standard Six

2026-07-25T15:29:17Z

Toro Andersen:

Studebaker Standard Six

Bore 85,72 mm

Stroke 115 mm

Volume 3982 ccm
[[File:Studebaker Standard Six.png|alt=(1925) Studebaker Standard Six|thumb|(1925) Studebaker Standard Six]]
Pr Cyl 663 ccm

File:Studebaker Standard Six.png

2026-07-25T15:27:33Z

Toro Andersen:

Studebaker Standard Six 1925

Studebaker Standard Six

2026-07-25T15:24:47Z

Toro Andersen: Created page with "Studebaker Standard Six"

Studebaker Standard Six

Hupmobile 8 - 1925

2026-07-25T08:16:26Z

Toro Andersen:

Hupmobile 8 cyl inline
[[File:Hupmobile 8 Cyl Inline.png|thumb|(1925) Hupmobile 8 Cyl Inline]]Bore 2 7/8 - 73,025

Stroke 4 3/4 - 120,62

Volume 4042 ccm

Pr Cyl 505 ccm

Weight 270 Kg

Ignition seq 1 - 5 - 2 - 3 - 8 - 4 - 7 - 6

Hupmobile 8 - 1925

2026-07-25T08:14:51Z

Toro Andersen:

Hupmobile 8 cyl inline
[[File:Hupmobile 8 Cyl Inline.png|thumb|(1925) Hupmobile 8 Cyl Inline]]Bore 2 7/8 - 73,025

Stroke 4 3/4 - 120,62

Volume 4042 ccm

Pr Cyl 505 ccm

Hupmobile 8 - 1925

2026-07-25T08:11:22Z

Toro Andersen:

Hupmobile 8 cyl inline
[[File:Hupmobile 8 Cyl Inline.png|thumb|(1925) Hupmobile 8 Cyl Inline]]

File:Hupmobile 8 Cyl Inline.png

2026-07-25T08:11:07Z

Toro Andersen:

Hupmobile 8 Cyl Inline

Hupmobile 8 - 1925

2026-07-25T08:10:31Z

Toro Andersen: Created page with "Hupmobile 8 cyl inline"

Hupmobile 8 cyl inline

Dodge L-Head 1925

2026-07-25T08:07:57Z

Toro Andersen:

Dodge L-Head 1925
[[File:Dodge L-Head 1925.png|alt=(1925) Dodge L-Head|thumb|(1925) Dodge L-Head]]

File:Dodge L-Head 1925.png

2026-07-25T08:07:39Z

Toro Andersen:

Dodge 4 cyl 1925

Dodge L-Head 1925

2026-07-25T08:07:05Z

Toro Andersen: Created page with "Dodge L-Head 1925"

Dodge L-Head 1925

Chevrolet 4 Cyl 1925

2026-07-25T08:02:00Z

Toro Andersen:

Chevrolet 4 Cyl engine 1925

Bore 93,8 mm

stroke 101 mm
[[File:Chevrolet 4 Cyl 1925.png|thumb|(1925) Chevrolet 1925]]
Volume 2792 ccm

Pr Cyl 697 ccm

File:Chevrolet 4 Cyl 1925.png

2026-07-25T07:59:04Z

Toro Andersen:

Chevrolet 4 Cyl 1925

2026-07-25T07:57:51Z

Toro Andersen: Created page with "Chevrolet 4 Cyl engine 1925"

Chevrolet 4 Cyl engine 1925

Nash Special Six

2026-07-25T07:55:40Z

Toro Andersen:

{{PAGENAME}}

[[File:Nash Special Six 1925.png|alt=(1925) Nash Special Six|thumb|(1925) Nash Special Six ]]

File:Nash Special Six 1925.png

2026-07-25T07:55:16Z

Toro Andersen:

Nash Special Six 1925

Nash Special Six

2026-07-25T07:54:03Z

Toro Andersen: Created page with "{{PAGENAME}}"

Nozzle and Injector list

2026-07-24T19:51:18Z

Toro Andersen: /* Detroit Diesel Injector Information */

List of old engines Nozzles and injectors, compiled by Toro

American Bosch 
Belarus-Kamaz 
Bendix 
Bosch 
Caterpillar 
Cummins 
Delphi 
Denso 
Detroit Diesel 
Doowon 
F&M 
Fiat - Iveco 
Ford-Navistar 
GMT 
L´Orange 
Lucas-Bryce 
Lucas-Simms 
Lucas-CAV 
MaK 
MAN 
Motorpal 
OMAP 
Perkins 
Ruston 
Seven Diesel 
Siemens VDO 
Sigma 
SKL 
Stanadyne 
Sulzer 
Yanmar 
Zexel 

{| class="wikitable"
|+
!
!
!
!
|-
|
|
|
|
|-
|
|
|
|
|-
|
|
|
|
|}

== Mercedes-Benz Nozzle ==
{| class="wikitable"
|+<big>Mercedes-Benz</big>
!Engine
!Variant
!Year
!OEM
!Stamp mark
!Bosch
!Lucas-CAV
!Monark
!Magneti Marelli
!Delphi
!Pressure
|-
|OM 312
|
|
|
|
|
|
|
|
|
|
|-
|OM 352
|
|
|000 017 39 12
|BDLLA 150 S 187
|
|5621 867
|
|
|
|200
|-
|OM 355
|
|
|
|BDLLA 150 S 186
|
|
|
|
|
|175
|-
|OM 615
|
|
|000 017 28 12
|BDN 0 SD 1510
|0 434 250 011
|5641 894
|39-305-011
|41 77462900 00
|
|115
|-
|OM 616
|
|
|
|
|
|
|
|
|
|
|-
|OM 617
|
|
|
|
|
|
|
|
|
|
|-
|OM 621
|
|1959 - 1961
|
|
|
|
|
|
|
|
|-
|OM 632
|
|
|
|
|
|
|
|
|
|
|-
|OM 636
|
|
|000 017 13 12
|DN 0 SD 211
|0 434 250 009
|LDFN0384
|
|
|
|
|}

== Detroit Diesel Injector Information ==
{| class="wikitable"
|+<big>Series 53 & 71 High Valve Configuration Injectors</big>
! rowspan="2" |Engine
! rowspan="2" |Injector
! rowspan="2" |Parts
Catalog
Type
! colspan="2" |Assy. Part
No
! rowspan="2" |Plunger
Design
! rowspan="2" |P & B
Part
No
! rowspan="2" |Spray
Tip
! rowspan="2" |Tip
Part No.
! rowspan="2" |Calibrator
Min / Max
|-
!Standard
Body
!Offset
Body
|-
| rowspan="5" |53
|35
|41
| -
|5228535[[#1| 1]]
|35
|5228530
| rowspan="2" |8- .005-165°
| rowspan="2" |5228436
|36-40
|-
|40
|47
| -
|5228560
| rowspan="2" |40
|5228552
|39-43
|-
|S40[[#2| 2]]
|51
| -
|5228590[[#1| 1]]
|5228552
| rowspan="3" |8- .055-165°
| rowspan="3" |5228473
|43-47
|-
|S45[[#2| 2]]
|52
| -
|5228580[[#1| 1]]
|45
|5228495
|48-52
|-
|S50[[#2| 2]]
|56
| -
|5228550[[#1| 1]]
|55
| rowspan="2" |5228126
|52-56
|-
| rowspan="7" |71
|HV55
|14
|5228300
|
|55
| rowspan="2" |6- .006-155° -H
| rowspan="2" |5228296
|52-56
|-
|HV6
|11
|5228310
|
|6H
|5228306
|59-63
|-
|HV7
|12. 25
|5228305
|5228470
|7H
|5228304
| rowspan="5" |7-.006-165°-H
| rowspan="5" |5228554
|71-75
|-
|HV8
|13. 24
|5228110
|5228435
|8H
|5228101
|80-84
|-
|HV9
|21
|5228380
| -
|9H
|5228378
|87-92
|-
|S80
|50
| -
|5228524
|8H
|5228101
|80-84
|-
|S90
|48
| -
|5228525[[#1| 1]]
|9X
|5228549
|87-92
|}

{| class="wikitable"
|+<big>SUMMARY OF SERIES 71 "A","B", and "E" INJECTORS</big>
!Injector
!Parts
Calaog
Type
!Assy
Part No.
!Plunger
Design
!P & B
Part No.
!Tip
part No.
!Spray Tip
!Calibrator
Min / Max
|-
|B55
|128
|5229555
|45C
|5229346
| rowspan="2" |5229532
|8-.006-162°[[#3| 3]]
|53-58
|-
|B60
|129
|5229560
|5C
|5229354
|8-.006-162°[[#3| 3]]
|57-62
|-
|B65
|130
|5229565
|65C
|5229542
|5229030
|8-.006-165°A
|64-69
|-
|71B5
|131
|5229550
|4C
|5229342
|5229532
|8-.006-162°[[#3| 3]]
|50-55
|-
|B55E
|132
|5229575
|B55E
|5229559
| rowspan="2" |5229553
|8-.006-162°[[#3| 3]]
|53-58
|-
|7B5E
|133
|5229580
|7B5E
|5229562
|8-.006-162°[[#3| 3]]
|50-55
|-
|7A50
|152
|5229690
|A5
|5229692
| rowspan="3" |5229642
| rowspan="3" |9-.0055-165A
|53-58
|-
|7A55
|151
|5229685
|A55
|5229686
|56-61
|-
|7A60
|148
|5229680
|A6
|5229644
|60-65
|-
|7A65
|150
|5229665
|65A
|5229664
| rowspan="3" |5229636
| rowspan="3" |9-.0058-165A
|66-71
|-
|7A70
|149
|5229670
|A7
|5229672
|72-77
|-
|7A75
|147
|5229675
|75A
|5229638
|75-80
|-
|7E50
|165
|5229785
|5E
|5229791
| rowspan="3" |5229026
| rowspan="3" |8-.0055-165°A
|50-55
|-
|7E55
|163
|5229790
|55E
|5229823
|53-58
|-
|7E60
|164
|5229795
|6E
|5229814
|57-62
|-
|7E65
|165
|5229800
|65E
|5229782
|5229030
|8-.006-165°A
|63-68
|-
|7E75
|171
|5229820
|N8
|5229808
|5229642
|9-.0055-165°A
|75-80
|}
*
* <div id="1">1 Low clamp bodies - Standard on Series 53 inline and V and All 4 Valve Series 71 engines.</div>
* <div id="2">2 Use only with low velocity camshafts on 3-53 (5126745-LH, 5126746-RH) 4-53 (5126747-LH, 5126748-RH) and 6V-53 (5126833-LH, 5126834-RH) </div>
* <div id="3">3 Designation not shown on spray tip.</div>

Nozzle and Injector list

2026-07-24T19:46:37Z

Toro Andersen:

List of old engines Nozzles and injectors, compiled by Toro

American Bosch 
Belarus-Kamaz 
Bendix 
Bosch 
Caterpillar 
Cummins 
Delphi 
Denso 
Detroit Diesel 
Doowon 
F&M 
Fiat - Iveco 
Ford-Navistar 
GMT 
L´Orange 
Lucas-Bryce 
Lucas-Simms 
Lucas-CAV 
MaK 
MAN 
Motorpal 
OMAP 
Perkins 
Ruston 
Seven Diesel 
Siemens VDO 
Sigma 
SKL 
Stanadyne 
Sulzer 
Yanmar 
Zexel 

{| class="wikitable"
|+
!
!
!
!
|-
|
|
|
|
|-
|
|
|
|
|-
|
|
|
|
|}

== Mercedes-Benz Nozzle ==
{| class="wikitable"
|+<big>Mercedes-Benz</big>
!Engine
!Variant
!Year
!OEM
!Stamp mark
!Bosch
!Lucas-CAV
!Monark
!Magneti Marelli
!Delphi
!Pressure
|-
|OM 312
|
|
|
|
|
|
|
|
|
|
|-
|OM 352
|
|
|000 017 39 12
|BDLLA 150 S 187
|
|5621 867
|
|
|
|200
|-
|OM 355
|
|
|
|BDLLA 150 S 186
|
|
|
|
|
|175
|-
|OM 615
|
|
|000 017 28 12
|BDN 0 SD 1510
|0 434 250 011
|5641 894
|39-305-011
|41 77462900 00
|
|115
|-
|OM 616
|
|
|
|
|
|
|
|
|
|
|-
|OM 617
|
|
|
|
|
|
|
|
|
|
|-
|OM 621
|
|1959 - 1961
|
|
|
|
|
|
|
|
|-
|OM 632
|
|
|
|
|
|
|
|
|
|
|-
|OM 636
|
|
|000 017 13 12
|DN 0 SD 211
|0 434 250 009
|LDFN0384
|
|
|
|
|}

== Detroit Diesel Injector Information ==
{| class="wikitable"
|+<big>Series 53 & 71 High Valve Configuration Injectors</big>
! rowspan="2" |Engine
! rowspan="2" |Injector
! rowspan="2" |Parts
Catalog
Type
! colspan="2" |Assy. Part
No
! rowspan="2" |Plunger
Design
! rowspan="2" |P & B
Part
No
! rowspan="2" |Spray
Tip
! rowspan="2" |Tip
Part No.
! rowspan="2" |Calibrator
Min / Max
|-
!Standard
Body
!Offset
Body
|-
| rowspan="5" |53
|35
|41
| -
|5228535[[#1|1]]
|35
|5228530
| rowspan="2" |8- .005-165°
| rowspan="2" |5228436
|36-40
|-
|40
|47
| -
|5228560
| rowspan="2" |40
|5228552
|39-43
|-
|S40[[#2|2]]
|51
| -
|5228590[[#1|1]]
|5228552
| rowspan="3" |8- .055-165°
| rowspan="3" |5228473
|43-47
|-
|S45[[#2|2]]
|52
| -
|5228580[[#1|1]]
|45
|5228495
|48-52
|-
|S50[[#2|2]]
|56
| -
|5228550[[#1|1]]
|55
| rowspan="2" |5228126
|52-56
|-
| rowspan="7" |71
|HV55
|14
|5228300
|
|55
| rowspan="2" |6- .006-155° -H
| rowspan="2" |5228296
|52-56
|-
|HV6
|11
|5228310
|
|6H
|5228306
|59-63
|-
|HV7
|12. 25
|5228305
|5228470
|7H
|5228304
| rowspan="5" |7-.006-165°-H
| rowspan="5" |5228554
|71-75
|-
|HV8
|13. 24
|5228110
|5228435
|8H
|5228101
|80-84
|-
|HV9
|21
|5228380
| -
|9H
|5228378
|87-92
|-
|S80
|50
| -
|5228524
|8H
|5228101
|80-84
|-
|S90
|48
| -
|5228525[[#1|1]]
|9X
|5228549
|87-92
|}

{| class="wikitable"
|+<big>SUMMARY OF SERIES 71 "A","B", and "E" INJECTORS</big>
!Injector
!Parts
Calaog
Type
!Assy
Part No.
!Plunger
Design
!P & B
Part No.
!Tip
part No.
!Spray Tip
!Calibrator
Min / Max
|-
|B55
|128
|5229555
|45C
|5229346
| rowspan="2" |5229532
|8-.006-162°[[#3|3]]
|53-58
|-
|B60
|129
|5229560
|5C
|5229354
|8-.006-162°[[#3|3]]
|57-62
|-
|B65
|130
|5229565
|65C
|5229542
|5229030
|8-.006-165°A
|64-69
|-
|71B5
|131
|5229550
|4C
|5229342
|5229532
|8-.006-162°[[#3|3]]
|50-55
|-
|B55E
|132
|5229575
|B55E
|5229559
| rowspan="2" |5229553
|8-.006-162°[[#3|3]]
|53-58
|-
|7B5E
|133
|5229580
|7B5E
|5229562
|8-.006-162°[[#3|3]]
|50-55
|-
|7A50
|152
|5229690
|A5
|5229692
| rowspan="3" |5229642
| rowspan="3" |9-.0055-165A
|53-58
|-
|7A55
|151
|5229685
|A55
|5229686
|56-61
|-
|7A60
|148
|5229680
|A6
|5229644
|60-65
|-
|7A65
|150
|5229665
|65A
|5229664
| rowspan="3" |5229636
| rowspan="3" |9-.0058-165A
|66-71
|-
|7A70
|149
|5229670
|A7
|5229672
|72-77
|-
|7A75
|147
|5229675
|75A
|5229638
|75-80
|-
|7E50
|165
|5229785
|5E
|5229791
| rowspan="3" |5229026
| rowspan="3" |8-.0055-165°A
|50-55
|-
|7E55
|163
|5229790
|55E
|5229823
|53-58
|-
|7E60
|164
|5229795
|6E
|5229814
|57-62
|-
|7E65
|165
|5229800
|65E
|5229782
|5229030
|8-.006-165°A
|63-68
|-
|7E75
|171
|5229820
|N8
|5229808
|5229642
|9-.0055-165°A
|75-80
|}
*
* <div id="1">1 Low clamp bodies - Standard on Series 53 inline and V and All 4 Valve Series 71 engines.</div>
* <div id="2">2 Use only with low velocity camshafts on 3-53 (5126745-LH, 5126746-RH) 4-53 (5126747-LH, 5126748-RH) and 6V-53 (5126833-LH, 5126834-RH) </div>
* <div id="3">3 Designation not shown on spray tip.</div>

Nozzle and Injector list

2026-07-24T19:43:45Z

Toro Andersen:

List of old engines Nozzles and injectors, compiled by Toro

American Bosch 
Belarus-Kamaz 
Bendix 
Bosch 
Caterpillar 
Cummins 
Delphi 
Denso 
Detroit Diesel 
Doowon 
F&M 
Fiat - Iveco 
Ford-Navistar 
GMT 
L´Orange 
Lucas-Bryce 
Lucas-Simms 
Lucas-CAV 
MaK 
MAN 
Motorpal 
OMAP 
Perkins 
Ruston 
Seven Diesel 
Siemens VDO 
Sigma 
SKL 
Stanadyne 
Sulzer 
Yanmar 
Zexel 

{| class="wikitable"
|+
!
!
!
!
|-
|
|
|
|
|-
|
|
|
|
|-
|
|
|
|
|}

== Mercedes-Benz Nozzle ==
{| class="wikitable"
|+<big>Mercedes-Benz</big>
!Engine
!Variant
!Year
!OEM
!Stamp mark
!Bosch
!Lucas-CAV
!Monark
!Magneti Marelli
!Delphi
!Pressure
|-
|OM 312
|
|
|
|
|
|
|
|
|
|
|-
|OM 352
|
|
|000 017 39 12
|BDLLA 150 S 187
|
|5621 867
|
|
|
|200
|-
|OM 355
|
|
|
|BDLLA 150 S 186
|
|
|
|
|
|175
|-
|OM 615
|
|
|000 017 28 12
|BDN 0 SD 1510
|0 434 250 011
|5641 894
|39-305-011
|41 77462900 00
|
|115
|-
|OM 616
|
|
|
|
|
|
|
|
|
|
|-
|OM 617
|
|
|
|
|
|
|
|
|
|
|-
|OM 621
|
|1959 - 1961
|
|
|
|
|
|
|
|
|-
|OM 632
|
|
|
|
|
|
|
|
|
|
|-
|OM 636
|
|
|000 017 13 12
|DN 0 SD 211
|0 434 250 009
|LDFN0384
|
|
|
|
|}

== Detroit Diesel Injector Information ==
{| class="wikitable"
|+<big>Series 53 & 71 High Valve Configuration Injectors</big>
! rowspan="2" |Engine
! rowspan="2" |Injector
! rowspan="2" |Parts
Catalog
Type
! colspan="2" |Assy. Part
No
! rowspan="2" |Plunger
Design
! rowspan="2" |P & B
Part
No
! rowspan="2" |Spray
Tip
! rowspan="2" |Tip
Part No.
! rowspan="2" |Calibrator
Min / Max
|-
!Standard
Body
!Offset
Body
|-
| rowspan="5" |53
|35
|41
| -
|5228535[[#1|1]]
|35
|5228530
| rowspan="2" |8- .005-165°
| rowspan="2" |5228436
|36-40
|-
|40
|47
| -
|5228560
| rowspan="2" |40
|5228552
|39-43
|-
|S40[[#2|2]]
|51
| -
|5228590[[#1|1]]
|5228552
| rowspan="3" |8- .055-165°
| rowspan="3" |5228473
|43-47
|-
|S45[[#2|2]]
|52
| -
|5228580[[#1|1]]
|45
|5228495
|48-52
|-
|S50[[#2|2]]
|56
| -
|5228550[[#1|1]]
|55
| rowspan="2" |5228126
|52-56
|-
| rowspan="7" |71
|HV55
|14
|5228300
|
|55
| rowspan="2" |6- .006-155° -H
| rowspan="2" |5228296
|52-56
|-
|HV6
|11
|5228310
|
|6H
|5228306
|59-63
|-
|HV7
|12. 25
|5228305
|5228470
|7H
|5228304
| rowspan="5" |7-.006-165°-H
| rowspan="5" |5228554
|71-75
|-
|HV8
|13. 24
|5228110
|5228435
|8H
|5228101
|80-84
|-
|HV9
|21
|5228380
| -
|9H
|5228378
|87-92
|-
|S80
|50
| -
|5228524
|8H
|5228101
|80-84
|-
|S90
|48
| -
|5228525[[#1|1]]
|9X
|5228549
|87-92
|}
* <div id="1">1 Low clamp bodies - Standard on Series 53 inline and V and All 4 Valve Series 71 engines.</div>
* <div id="2">2 Use only with low velocity camshafts on 3-53 (5126745-LH, 5126746-RH) 4-53 (5126747-LH, 5126748-RH) and 6V-53 (5126833-LH, 5126834-RH) </div>
{| class="wikitable"
|+<big>SUMMARY OF SERIES 71 "A","B", and "E" INJECTORS</big>
!Injector
!Parts
Calaog
Type
!Assy
Part No.
!Plunger
Design
!P & B
Part No.
!Tip
part No.
!Spray Tip
!Calibrator
Min / Max
|-
|B55
|128
|5229555
|45C
|5229346
| rowspan="2" |5229532
|8-.006-162°
|53-58
|-
|B60
|129
|5229560
|5C
|5229354
|8-.006-162°
|57-62
|-
|B65
|130
|5229565
|65C
|5229542
|5229030
|8-.006-165°A
|64-69
|-
|71B5
|131
|5229550
|4C
|5229342
|5229532
|8-.006-162°
|50-55
|-
|B55E
|132
|5229575
|B55E
|5229559
| rowspan="2" |5229553
|8-.006-162°
|53-58
|-
|7B5E
|133
|5229580
|7B5E
|5229562
|8-.006-162°
|50-55
|-
|7A50
|152
|5229690
|A5
|5229692
| rowspan="3" |5229642
| rowspan="3" |9-.0055-165A
|53-58
|-
|7A55
|151
|5229685
|A55
|5229686
|56-61
|-
|7A60
|148
|5229680
|A6
|5229644
|60-65
|-
|7A65
|150
|5229665
|65A
|5229664
| rowspan="3" |5229636
| rowspan="3" |9-.0058-165A
|66-71
|-
|7A70
|149
|5229670
|A7
|5229672
|72-77
|-
|7A75
|147
|5229675
|75A
|5229638
|75-80
|-
|7E50
|165
|5229785
|5E
|5229791
| rowspan="3" |5229026
| rowspan="3" |8-.0055-165°A
|50-55
|-
|7E55
|163
|5229790
|55E
|5229823
|53-58
|-
|7E60
|164
|5229795
|6E
|5229814
|57-62
|-
|7E65
|165
|5229800
|65E
|5229782
|5229030
|8-.006-165°A
|63-68
|-
|7E75
|171
|5229820
|N8
|5229808
|5229642
|9-.0055-165°A
|75-80
|}
*

Nozzle and Injector list

2026-07-24T19:42:13Z

Toro Andersen: /* Detroit Diesel Injector Information */

List of old engines Nozzles and injectors, compiled by Toro

American Bosch 
Belarus-Kamaz 
Bendix 
Bosch 
Caterpillar 
Cummins 
Delphi 
Denso 
Detroit Diesel 
Doowon 
F&M 
Fiat - Iveco 
Ford-Navistar 
GMT 
L´Orange 
Lucas-Bryce 
Lucas-Simms 
Lucas-CAV 
MaK 
MAN 
Motorpal 
OMAP 
Perkins 
Ruston 
Seven Diesel 
Siemens VDO 
Sigma 
SKL 
Stanadyne 
Sulzer 
Yanmar 
Zexel 

{| class="wikitable"
|+
!
!
!
!
|-
|
|
|
|
|-
|
|
|
|
|-
|
|
|
|
|}
{| class="wikitable"
|+<big>Mercedes-Benz</big>
!Engine
!Variant
!Year
!OEM
!Stamp mark
!Bosch
!Lucas-CAV
!Monark
!Magneti Marelli
!Delphi
!Pressure
|-
|OM 312
|
|
|
|
|
|
|
|
|
|
|-
|OM 352
|
|
|000 017 39 12
|BDLLA 150 S 187
|
|5621 867
|
|
|
|200
|-
|OM 355
|
|
|
|BDLLA 150 S 186
|
|
|
|
|
|175
|-
|OM 615
|
|
|000 017 28 12
|BDN 0 SD 1510
|0 434 250 011
|5641 894
|39-305-011
|41 77462900 00
|
|115
|-
|OM 616
|
|
|
|
|
|
|
|
|
|
|-
|OM 617
|
|
|
|
|
|
|
|
|
|
|-
|OM 621
|
|1959 - 1961
|
|
|
|
|
|
|
|
|-
|OM 632
|
|
|
|
|
|
|
|
|
|
|-
|OM 636
|
|
|000 017 13 12
|DN 0 SD 211
|0 434 250 009
|LDFN0384
|
|
|
|
|}

== Detroit Diesel Injector Information ==
{| class="wikitable"
|+<big>Series 53 & 71 High Valve Configuration Injectors</big>
! rowspan="2" |Engine
! rowspan="2" |Injector
! rowspan="2" |Parts
Catalog
Type
! colspan="2" |Assy. Part
No
! rowspan="2" |Plunger
Design
! rowspan="2" |P & B
Part
No
! rowspan="2" |Spray
Tip
! rowspan="2" |Tip
Part No.
! rowspan="2" |Calibrator
Min / Max
|-
!Standard
Body
!Offset
Body
|-
| rowspan="5" |53
|35
|41
| -
|5228535[[#1|1]]
|35
|5228530
| rowspan="2" |8- .005-165°
| rowspan="2" |5228436
|36-40
|-
|40
|47
| -
|5228560
| rowspan="2" |40
|5228552
|39-43
|-
|S40[[#2|2]]
|51
| -
|5228590[[#1|1]]
|5228552
| rowspan="3" |8- .055-165°
| rowspan="3" |5228473
|43-47
|-
|S45[[#2|2]]
|52
| -
|5228580[[#1|1]]
|45
|5228495
|48-52
|-
|S50[[#2|2]]
|56
| -
|5228550[[#1|1]]
|55
| rowspan="2" |5228126
|52-56
|-
| rowspan="7" |71
|HV55
|14
|5228300
|
|55
| rowspan="2" |6- .006-155° -H
| rowspan="2" |5228296
|52-56
|-
|HV6
|11
|5228310
|
|6H
|5228306
|59-63
|-
|HV7
|12. 25
|5228305
|5228470
|7H
|5228304
| rowspan="5" |7-.006-165°-H
| rowspan="5" |5228554
|71-75
|-
|HV8
|13. 24
|5228110
|5228435
|8H
|5228101
|80-84
|-
|HV9
|21
|5228380
| -
|9H
|5228378
|87-92
|-
|S80
|50
| -
|5228524
|8H
|5228101
|80-84
|-
|S90
|48
| -
|5228525[[#1|1]]
|9X
|5228549
|87-92
|}
* <div id="1">1 Low clamp bodies - Standard on Series 53 inline and V and All 4 Valve Series 71 engines.</div>
* <div id="2">2 Use only with low velocity camshafts on 3-53 (5126745-LH, 5126746-RH) 4-53 (5126747-LH, 5126748-RH) and 6V-53 (5126833-LH, 5126834-RH) </div>
{| class="wikitable"
|+<big>SUMMARY OF SERIES 71 "A","B", and "E" INJECTORS</big>
!Injector
!Parts
Calaog
Type
!Assy
Part No.
!Plunger
Design
!P & B
Part No.
!Tip
part No.
!Spray Tip
!Calibrator
Min / Max
|-
|B55
|128
|5229555
|45C
|5229346
| rowspan="2" |5229532
|8-.006-162°
|53-58
|-
|B60
|129
|5229560
|5C
|5229354
|8-.006-162°
|57-62
|-
|B65
|130
|5229565
|65C
|5229542
|5229030
|8-.006-165°A
|64-69
|-
|71B5
|131
|5229550
|4C
|5229342
|5229532
|8-.006-162°
|50-55
|-
|B55E
|132
|5229575
|B55E
|5229559
| rowspan="2" |5229553
|8-.006-162°
|53-58
|-
|7B5E
|133
|5229580
|7B5E
|5229562
|8-.006-162°
|50-55
|-
|7A50
|152
|5229690
|A5
|5229692
| rowspan="3" |5229642
| rowspan="3" |9-.0055-165A
|53-58
|-
|7A55
|151
|5229685
|A55
|5229686
|56-61
|-
|7A60
|148
|5229680
|A6
|5229644
|60-65
|-
|7A65
|150
|5229665
|65A
|5229664
| rowspan="3" |5229636
| rowspan="3" |9-.0058-165A
|66-71
|-
|7A70
|149
|5229670
|A7
|5229672
|72-77
|-
|7A75
|147
|5229675
|75A
|5229638
|75-80
|-
|7E50
|165
|5229785
|5E
|5229791
| rowspan="3" |5229026
| rowspan="3" |8-.0055-165°A
|50-55
|-
|7E55
|163
|5229790
|55E
|5229823
|53-58
|-
|7E60
|164
|5229795
|6E
|5229814
|57-62
|-
|7E65
|165
|5229800
|65E
|5229782
|5229030
|8-.006-165°A
|63-68
|-
|7E75
|171
|5229820
|N8
|5229808
|5229642
|9-.0055-165°A
|75-80
|}
*

Nozzle and Injector list

2026-07-24T16:38:43Z

Toro Andersen:

List of old engines Nozzles and injectors, compiled by Toro

American Bosch 
Belarus-Kamaz 
Bendix 
Bosch 
Caterpillar 
Cummins 
Delphi 
Denso 
Detroit Diesel 
Doowon 
F&M 
Fiat - Iveco 
Ford-Navistar 
GMT 
L´Orange 
Lucas-Bryce 
Lucas-Simms 
Lucas-CAV 
MaK 
MAN 
Motorpal 
OMAP 
Perkins 
Ruston 
Seven Diesel 
Siemens VDO 
Sigma 
SKL 
Stanadyne 
Sulzer 
Yanmar 
Zexel 

{| class="wikitable"
|+
!
!
!
!
|-
|
|
|
|
|-
|
|
|
|
|-
|
|
|
|
|}
{| class="wikitable"
|+<big>Mercedes-Benz</big>
!Engine
!Variant
!Year
!OEM
!Stamp mark
!Bosch
!Lucas-CAV
!Monark
!Magneti Marelli
!Delphi
!Pressure
|-
|OM 312
|
|
|
|
|
|
|
|
|
|
|-
|OM 352
|
|
|000 017 39 12
|BDLLA 150 S 187
|
|5621 867
|
|
|
|200
|-
|OM 355
|
|
|
|BDLLA 150 S 186
|
|
|
|
|
|175
|-
|OM 615
|
|
|000 017 28 12
|BDN 0 SD 1510
|0 434 250 011
|5641 894
|39-305-011
|41 77462900 00
|
|115
|-
|OM 616
|
|
|
|
|
|
|
|
|
|
|-
|OM 617
|
|
|
|
|
|
|
|
|
|
|-
|OM 621
|
|1959 - 1961
|
|
|
|
|
|
|
|
|-
|OM 632
|
|
|
|
|
|
|
|
|
|
|-
|OM 636
|
|
|000 017 13 12
|DN 0 SD 211
|0 434 250 009
|LDFN0384
|
|
|
|
|}

== Detroit Diesel Injector Information ==
{| class="wikitable"
|+<big>Series 53 & 71 High Valve Configuration Injectors</big>
! rowspan="2" |Engine
! rowspan="2" |Injector
! rowspan="2" |Parts
Catalog
Type
! colspan="2" |Assy. Part
No
! rowspan="2" |Plunger
Design
! rowspan="2" |P & B
Part
No
! rowspan="2" |Spray
Tip
! rowspan="2" |Tip
Part No.
! rowspan="2" |Calibrator
Min / Max
|-
!Standard
Body
!Offset
Body
|-
| rowspan="5" |53
|35
|41
| -
|5228535[[#1|1]]
|35
|5228530
| rowspan="2" |8- .005-165°
| rowspan="2" |5228436
|36-40
|-
|40
|47
| -
|5228560
| rowspan="2" |40
|5228552
|39-43
|-
|S40[[#2|2]]
|51
| -
|5228590[[#1|1]]
|5228552
| rowspan="3" |8- .055-165°
| rowspan="3" |5228473
|43-47
|-
|S45[[#2|2]]
|52
| -
|5228580[[#1|1]]
|45
|5228495
|48-52
|-
|S50[[#2|2]]
|56
| -
|5228550[[#1|1]]
|55
| rowspan="2" |5228126
|52-56
|-
| rowspan="7" |71
|HV55
|14
|5228300
|
|55
| rowspan="2" |6- .006-155° -H
| rowspan="2" |5228296
|52-56
|-
|HV6
|11
|5228310
|
|6H
|5228306
|59-63
|-
|HV7
|12. 25
|5228305
|5228470
|7H
|5228304
| rowspan="5" |7-.006-165°-H
| rowspan="5" |5228554
|71-75
|-
|HV8
|13. 24
|5228110
|5228435
|8H
|5228101
|80-84
|-
|HV9
|21
|5228380
| -
|9H
|5228378
|87-92
|-
|S80
|50
| -
|5228524
|8H
|5228101
|80-84
|-
|S90
|48
| -
|5228525[[#1|1]]
|9X
|5228549
|87-92
|}
* <div id="1">1 Low clamp bodies - Standard on Series 53 inline and V and All 4 Valve Series 71 engines.</div>
* <div id="2">2 Use only with low velocity camshafts on 3-53 (5126745-LH, 5126746-RH) 4-53 (5126747-LH, 5126748-RH) and 6V-53 (5126833-LH, 5126834-RH) </div>
{| class="wikitable"
|+<big>SUMMARY OF SERIES 71 "A","B", and "E" INJECTORS</big>
!Injector
!Parts
Calaog
Type
!Assy
Part No.
!Plunger
Design
!P & B
Part No.
!Tip
part No.
!Spray Tip
!Calibrator
Min / Max
|-
|B55
|128
|5229555
|45C
|
| rowspan="2" |5229532
|8-.006-162°
|53-58
|-
|B60
|129
|5229560
|5C
|
|8-.006-162°
|57-62
|-
|B65
|130
|5229565
|65C
|
|
|8-.006-165°A
|64-69
|-
|71B5
|131
|5229550
|4C
|
|
|8-.006-162°
|50-55
|-
|B55E
|132
|5229575
|B55E
|
|
|8-.006-162°
|53-58
|-
|7B5E
|133
|5229580
|7B5E
|
|
|8-.006-162°
|50-55
|-
|7A50
|152
|5229690
|A5
|
|
|
|53-58
|-
|7A55
|151
|5229685
|A55
|
|
|
|56-61
|-
|7A60
|148
|
|A6
|
|
|
|60-65
|-
|7A65
|150
|
|65A
|
|
|
|66-71
|-
|7A70
|149
|
|A7
|
|
|
|72-77
|-
|7A75
|147
|
|75A
|
|
|
|75-80
|-
|7E50
|165
|
|5E
|
|
|
|50-55
|-
|7E55
|163
|
|55E
|
|
|
|53-58
|-
|7E60
|164
|
|6E
|
|
|
|57-62
|-
|7E65
|165
|
|65E
|
|
|
|63-68
|-
|7E75
|171
|
|N8
|
|
|
|75-80
|}
*

Nozzle and Injector list

2026-07-24T16:37:39Z

Toro Andersen: /* Detroit Diesel Injector Information */

List of old engines Nozzles and injectors, compiled by Toro

American Bosch 
Belarus-Kamaz 
Bendix 
Bosch 
Caterpillar 
Cummins 
Delphi 
Denso 
Detroit Diesel 
Doowon 
F&M 
Fiat - Iveco 
Ford-Navistar 
GMT 
L´Orange 
Lucas-Bryce 
Lucas-Simms 
Lucas-CAV 
MaK 
MAN 
Motorpal 
OMAP 
Perkins 
Ruston 
Seven Diesel 
Siemens VDO 
Sigma 
SKL 
Stanadyne 
Sulzer 
Yanmar 
Zexel 

{| class="wikitable"
|+
!
!
!
!
|-
|
|
|
|
|-
|
|
|
|
|-
|
|
|
|
|}
{| class="wikitable"
|+Mercedes-Benz
!Engine
!Variant
!Year
!OEM
!Stamp mark
!Bosch
!Lucas-CAV
!Monark
!Magneti Marelli
!Delphi
!Pressure
|-
|OM 312
|
|
|
|
|
|
|
|
|
|
|-
|OM 352
|
|
|000 017 39 12
|BDLLA 150 S 187
|
|5621 867
|
|
|
|200
|-
|OM 355
|
|
|
|BDLLA 150 S 186
|
|
|
|
|
|175
|-
|OM 615
|
|
|000 017 28 12
|BDN 0 SD 1510
|0 434 250 011
|5641 894
|39-305-011
|41 77462900 00
|
|115
|-
|OM 616
|
|
|
|
|
|
|
|
|
|
|-
|OM 617
|
|
|
|
|
|
|
|
|
|
|-
|OM 621
|
|1959 - 1961
|
|
|
|
|
|
|
|
|-
|OM 632
|
|
|
|
|
|
|
|
|
|
|-
|OM 636
|
|
|000 017 13 12
|DN 0 SD 211
|0 434 250 009
|LDFN0384
|
|
|
|
|}

== Detroit Diesel Injector Information ==
{| class="wikitable"
|+Series 53 & 71 High Valve Configuration Injectors
! rowspan="2" |Engine
! rowspan="2" |Injector
! rowspan="2" |Parts
Catalog
Type
! colspan="2" |Assy. Part
No
! rowspan="2" |Plunger
Design
! rowspan="2" |P & B
Part
No
! rowspan="2" |Spray
Tip
! rowspan="2" |Tip
Part No.
! rowspan="2" |Calibrator
Min / Max
|-
!Standard
Body
!Offset
Body
|-
| rowspan="5" |53
|35
|41
| -
|5228535[[#1|1]]
|35
|5228530
| rowspan="2" |8- .005-165°
| rowspan="2" |5228436
|36-40
|-
|40
|47
| -
|5228560
| rowspan="2" |40
|5228552
|39-43
|-
|S40[[#2|2]]
|51
| -
|5228590[[#1|1]]
|5228552
| rowspan="3" |8- .055-165°
| rowspan="3" |5228473
|43-47
|-
|S45[[#2|2]]
|52
| -
|5228580[[#1|1]]
|45
|5228495
|48-52
|-
|S50[[#2|2]]
|56
| -
|5228550[[#1|1]]
|55
| rowspan="2" |5228126
|52-56
|-
| rowspan="7" |71
|HV55
|14
|5228300
|
|55
| rowspan="2" |6- .006-155° -H
| rowspan="2" |5228296
|52-56
|-
|HV6
|11
|5228310
|
|6H
|5228306
|59-63
|-
|HV7
|12. 25
|5228305
|5228470
|7H
|5228304
| rowspan="5" |7-.006-165°-H
| rowspan="5" |5228554
|71-75
|-
|HV8
|13. 24
|5228110
|5228435
|8H
|5228101
|80-84
|-
|HV9
|21
|5228380
| -
|9H
|5228378
|87-92
|-
|S80
|50
| -
|5228524
|8H
|5228101
|80-84
|-
|S90
|48
| -
|5228525[[#1|1]]
|9X
|5228549
|87-92
|}
* <div id="1">1 Low clamp bodies - Standard on Series 53 inline and V and All 4 Valve Series 71 engines.</div>
* <div id="2">2 Use only with low velocity camshafts on 3-53 (5126745-LH, 5126746-RH) 4-53 (5126747-LH, 5126748-RH) and 6V-53 (5126833-LH, 5126834-RH) </div>
{| class="wikitable"
|+SUMMARY OF SERIES 71 A,"B, and E INJECTORS
!Injector
!Parts
Calaog
Type
!Assy
Part No.
!Plunger
Design
!P & B
Part No.
!Tip
part No.
!Spray Tip
!Calibrator
Min / Max
|-
|B55
|128
|5229555
|45C
|
| rowspan="2" |5229532
|8-.006-162°
|53-58
|-
|B60
|129
|5229560
|5C
|
|8-.006-162°
|57-62
|-
|B65
|130
|5229565
|65C
|
|
|8-.006-165°A
|64-69
|-
|71B5
|131
|5229550
|4C
|
|
|8-.006-162°
|50-55
|-
|B55E
|132
|5229575
|B55E
|
|
|8-.006-162°
|53-58
|-
|7B5E
|133
|5229580
|7B5E
|
|
|8-.006-162°
|50-55
|-
|7A50
|152
|5229690
|A5
|
|
|
|53-58
|-
|7A55
|151
|5229685
|A55
|
|
|
|56-61
|-
|7A60
|148
|
|A6
|
|
|
|60-65
|-
|7A65
|150
|
|65A
|
|
|
|66-71
|-
|7A70
|149
|
|A7
|
|
|
|72-77
|-
|7A75
|147
|
|75A
|
|
|
|75-80
|-
|7E50
|165
|
|5E
|
|
|
|50-55
|-
|7E55
|163
|
|55E
|
|
|
|53-58
|-
|7E60
|164
|
|6E
|
|
|
|57-62
|-
|7E65
|165
|
|65E
|
|
|
|63-68
|-
|7E75
|171
|
|N8
|
|
|
|75-80
|}
*