Internal combustion engine

It is disclosed a boxer engine with two substantially mirror-symmetric engine sides (L, R) comprising a crankshaft (1) to which is connected,

CROSS-REFERENCE TO RELATED APPLICATIONS

This patent is a U.S. national stage application of International Patent Application No. PCT/EP2018/086354 which was filed on Dec. 20, 2018 under the Patent Cooperation Treaty (PCT), which claims priority to European Patent Application No. 18153629.3 which was filed on Jan. 26, 2018, all of the foregoing applications are hereby incorporated herein by reference in their entireties.

TECHNICAL FIELD

The present invention relates generally to an internal combustion engine, in particular an internal combustion engine with low emission, for use in automobiles.

BACKGROUND

Ever since the internal combustion engine was first introduced centuries ago, it has continuously been developed and modified in order to adapt to the ever-changing demands in the market. Recent trends are increasingly concerned with environmental aspects and a sustainable future, calling for engines with lower emissions, which at this point can only be achieved by lowering the fuel consumption. Some of the concepts that have been introduced, with the intention of lowering the fuel consumption, are split cycle processes, variable valve timing and variable compression ratio.

A split cycle process occurs when the compression or expansion, or both, takes place in two or several stages. In theory, this concept should provide increased efficiency, but verification testing has shown increased mechanical and thermal losses, yielding insufficient payback for its complexity, additional weight and increased production cost.

In spark ignited engines, with a constant compression ratio, which use suction throttles for controlling the output power, a reduction of the filling ratio will cause a reduced pressure at the end of a compression stroke. Hence, the efficiency factor will decrease as the filling ratio decreases. To maintain a stable efficiency factor, thus increasing its overall efficiency, the compression ratio must be adjusted according to the filling ratio. Variable compression engines allow for the volume above the piston at top dead centre (TDC) to be changed. For automotive use, this needs to be done dynamically in response to the load and driving demands, as higher loads require lower ratios to be more efficient and vice versa. However, also this concept requires complex and heavy mechanisms, causing high production costs. This concept has also faced issues with vibrations. An example of prior art is disclosed by EP1170482.

Variable valve timing, also known as variable valve lift (used by Nissan) or “variable onckenwellen steuerung” (used by BMW, Ford, Ferrari and Lamborghini), makes it possible to adjust the opening times (lift, duration or both) for the suction or exhaust side valves whilst the engine is in operation. Variable valve timing can provide the benefits of internal exhaust gas recirculation, increased torque and better fuel economy, but production is expensive.

Another concept with beneficial features is the scotch yoke principle. Some of the features are exact sinusoidal reciprocating parts, fully dynamic mass balance which makes it vibration free, and options for simple double acting piston arrangements. Scotch yoke mechanisms are widely used in piston pumps, valve actuators, sewing machines and engines, as seen in US2012272758.

SUMMARY OF THE INVENTION

The present invention has the objective of providing an internal combustion engine incorporating the above-mentioned concepts, which solves the identified disadvantages in order to reduce the emission.

Said objectives are fully or partially achieved by an engine according to the independent claims. Preferred embodiments are set forth in the dependent claims.

According to a first aspect, the invention relates to a boxer engine with two substantially mirror-symmetric engine sides comprising a crankshaft to which is connected, at least two main scotch yoke assemblies each having one main piston arranged inside one main cylinder of each engine side, and at least one auxiliary scotch yoke assembly having a pair of auxiliary pistons arranged inside a pair of auxiliary cylinders of each engine side, wherein the main scotch yoke assemblies are arranged synchronized on the crankshaft and the at least one auxiliary scotch yoke assembly is arranged 180° offset on the crankshaft, each auxiliary piston defining an outer space and an inner space within each auxiliary cylinder, the inner space facing the opposite engine side, wherein, said inner spaces of each auxiliary cylinder pair are in fluid communication and forming a compression chamber, said compression chamber comprises first and second check valves, wherein the auxiliary cylinder pair is adapted to suck in ambient air through the first check valve and compress and pump said air out through the second check valve into a main cylinder of the opposite engine side, and said outer spaces of each auxiliary cylinder pair are in fluid communication and are receiving pressurized exhaust gas from a main cylinder of the same engine side.

The advantage of such an engine is that it enables two split cycle processes to take place, i.e. a compression process and an expansion process. For the expansion process, rather than discharging the remaining pressure within a main cylinder after a complete expansion stroke, the remaining pressure in all main cylinders are transferred to an outer space of a corresponding auxiliary cylinder pair so it can be used to further power the crankshaft and/or the compression process; thus, increasing the efficiency factor of the engine which in turn contributes to reduced emissions. For the compression process, rather than starting a compression stroke with a main cylinder filled with air at atmospheric pressure, a compression stroke starts with a main cylinder filled with compressed air; thus, reducing the fuel consumption and emissions.

Another advantage of such an engine is that the linear motion of the reciprocating scotch yoke assembly contributes to reduce vibrations in the engine. The scotch yoke also makes the pistons centric stable.

According to an embodiment of the present invention, the auxiliary pistons comprise circumferentially arranged pressure trap grooves. Since the pistons are centric stable, replacing pistons rings with pressure trap grooves will significantly reduce the friction between the auxiliary pistons and the auxiliary cylinder liners. This friction reduction is an improvement with regard to mechanical loss.

According to a second aspect, the present invention relates to a boxer engine wherein each main scotch yoke assembly comprises a main piston rod with a polygonal cross-section for each engine side, wherein each main piston rod: at a first end has a swivel connection to the corresponding main piston; at a second end has a threaded connection to a stud projecting from a corresponding main yoke; and is embraced by a longitudinally sliding worm gear.

With this mechanism, it is achieved a robust and accurate adjustment of the compression ratio of the main cylinders, whilst at the same time having an uncomplicated design, which is an improvement with regards to weight and production cost.

According to an embodiment of the present invention, worm control shafts engage the worm gears of the same engine side, said worm control shafts being adjusted by means of hydraulic or electric actuators. In this way, the compression ratio of two main cylinders are simultaneously operated by one control shaft, which increases its precision, and by incorporating hydraulic or electric actuators, the precision is further increased.

According to a third aspect, the invention relates to a boxer engine comprising two connecting shafts connecting the crankshaft and the camshafts operating the suction valves and the discharge valves of the main cylinders and the exhaust valves of the auxiliary cylinders, wherein each connecting shaft: at a first end portion comprises first internal helical splines engaged with first external helical splines of a first protruding spindle of a first connecting shaft bevel gear, said first connecting shaft bevel gear being engaged with a cam shaft bevel gear connected to the camshaft; at a second end portion comprises second internal helical splines engaged with second external helical splines of a second protruding spindle of a second connecting shaft bevel gear, said second connecting shaft bevel gear being engaged with a crankshaft gear connected to the crankshaft; and has a length which allows some longitudinal movement of the connecting shaft along the first and second protruding spindles, wherein the first external helical splines and the second external helical splines are opposite threaded, and the first internal helical splines and the second internal helical splines are opposite threaded.

With this mechanism, it is achieved a robust and accurate adjustment of the valve timing, whilst at the same time having an uncomplicated design, which is an improvement with regard to weight and production cost.

According to an embodiment of the present invention, the connecting shafts are longitudinally adjusted simultaneously by means of hydraulic or electric actuators. In this way, the precision is increased.

According to another embodiment of the present invention, the boxer engine comprises a cam shaft with a double cam in a middle region. The double cam enables one camshaft to operate both the auxiliary cylinder pair and the two main cylinders of the same engine side, ref. table 1.

The main cylinders and the outer spaces of an auxiliary cylinder pair of the same engine side are preferably connected by a valve seat plate to facilitate the split cycle expansion process.

The compression chambers and the main cylinders are preferably connected by at least one connecting channel to facilitate the split cycle compression process. By making the connecting channel air cooled, the charge of air supplied to the main cylinders will be further compressed, which will reduce the fuel consumption and emissions.

Balancing the weight of the at least one auxiliary yoke assembly with the weight of the at least two main yokes assemblies will reduce vibrations in the engine, which will enhance its durability and performance.

A cylinder bottom plate sealing around the reciprocating auxiliary piston rod makes the compression chamber substantially air tight, which enables the split cycle compression process.

DETAILED DESCRIPTION

In the disclosed figures, there are illustrated a boxer type internal combustion engine.FIG. 1shows an isometric view of the assembled engine. The engine is divided into two engine sides R, L. which are defined by a plane P of symmetry, wherein the two engine sides R, L substantially are mirror images of each other. The engine of the present invention could be used as a mono side design. A mono side design would need an accumulator for the first stage compressed charge, and because of pulsation in this it would perform with a lower efficiency. Hence, the dual side design is preferred.

Scotch Yoke Mechanism

In the engine, the linear motion of the pistons7,8moving inside the cylinders are converted into rotational motion of the crankshaft1, by the scotch yoke assemblies110,120. As detailed inFIG. 4andFIG. 5, the engine has two types of scotch yoke assemblies110,120, respectively a main scotch yoke assembly110and an auxiliary scotch yoke assembly120.FIG. 2shows a setup with a middle auxiliary scotch yoke assembly120and two outer main scotch yoke assemblies110.

The main scotch yoke assemblies110comprise a main yoke2, two crankshaft bearing halves6, two studs25, two main piston rods5and two main pistons7. The main pistons7are connected to the main piston rods5with swivel couplings28, illustrated inFIG. 4detail b. The main piston7has a slot in the swivel coupling28, permitting the main piston7to be assembled sideways onto the main piston rod5. This type of coupling will allow the main piston rod5to rotate freely relative to the main piston7. The main piston rod5has a swivel coupling28in a first end and internal threads27in a second end. The main piston rod5has a polygonal cross section. The studs25connect the main piston rods5to the main yoke2. The studs25can be attached to the main yoke2by means of welded or threaded connections, alternatively they can also be machined from the same piece. The main yoke2is substantially rectangular with sliding surfaces23fully or partly covering the upper and lower surfaces. The main piston rods5are positioned in central areas of the two side surfaces of the main yoke5, and are of the same length. The main yoke has a rectangular aperture in which the crankshaft bearing halves6are fitted. The crank shaft bearing halves6embrace the camshaft1. The two crankshaft halves6combined are adapted to a sliding motion in the longitudinal direction of the aperture.

The auxiliary scotch yoke assembly120comprises an auxiliary yoke3, two crankshaft bearing halves6, two auxiliary piston rods4and four auxiliary pistons8. The auxiliary pistons8are connected to the auxiliary piston rods4with a threaded and/or bolted connection. The auxiliary piston rods4are connected to the auxiliary yoke3with a bolted connection. The auxiliary yoke3is substantially rectangular, and has an aperture equal to the one of the main yoke2. Equal crankshaft bearing halves6are used in the auxiliary scotch yoke assembly120as in the main scotch yoke assembly110. Each auxiliary piston rod4has one auxiliary piston8connected to each of its two ends. Two auxiliary piston rods4are connected to the upper and lower surfaces of the auxiliary yoke3. Both auxiliary piston rods4protrudes an equal distance at both sides of the auxiliary yoke3, and both auxiliary piston rods4are of the same length. This means that the two auxiliary pistons8of a first engine side R, L will reach the top dead centre (TDC) simultaneously with the two auxiliary pistons8of a second engine side R, L reaching the bottom dead centre (BDC), and vice versa. Instead of piston rings, the auxiliary pistons8are equipped with pressure trap grooves72.

The weight of the auxiliary scotch yoke assembly120is balanced equal to the combined weight of the two main scotch yoke assemblies110. This is typically achieved by material selection, choosing materials with the desired mechanical properties, but with different density, e.g. steel and aluminium.

FIG. 3shows the same three scotch yoke assemblies110,120asFIG. 2. The scotch yoke assemblies110,120are arranged in guiding grooves77in an upper guiding plate50and a lower guiding plate51, which are mounted to a rear crankshaft bearing plate59.

Variable Compression Ratio

FIG. 3illustrates the mechanism enabling variable compression. By altering the top dead centre (TDC) of the main pistons7, a relatively constant compression pressure over the whole speed and load range can be achieved, i.e. the engine compression end pressure will remain on its decided value whatever the degree of charge filling in the main cylinders I, III; II, IV is. The variable compression mechanism of the present invention utilizes worm gears13,14and worm gear control shafts11,12to adjust the TDC of the main pistons7.

Worm gears13,14with a central polygonal aperture, corresponding to the cross section of the main piston rods5, are arranged on the main piston rods5. The worm gears13,14are adapted to rotate the main piston rods5, whilst the piston rods5can freely slide relative to the worm gears13,14in their longitudinal direction. As the worm gear13,14turns, the main piston rod5will travel the threads of the stud5. Since the stud5is static relative to the main yoke2, the travel of the main piston rod5will change its distance to the main yoke2. This will in turn change the distance between the main piston7and the corresponding main yoke2. When changing the distance between the main yoke2and the main piston7, the TDC of the same main piston7will be changed at an equal ratio.

A worm control shaft11,12is arranged on each engine side R, L, and kept in place by a cylinder bottom plate52. Each worm control shaft11,12has a worm in engagement with each worm gear13,14of the same engine side R, L, in this case two. The worm gears13,14and the worm control shafts11,12of opposite engine sides R, L are preferably made with opposite gears, e.g. the worm gears14of the left engine side L having left hand helical gears and the worm gears13of the right engine side R having right hand helical gears. In this way the TDC of the main pistons7on both engine sides R, L will change correspondingly when the worm control shafts11,12are rotated in the same direction, e.g. by turning both worm control shafts11,12clockwise, the TDC of all main pistons will be lowered. The worm control shafts11,12might be driven by means of hydraulic or electric actuators. Preferably the worm gear transmission has a high reduction ratio. One of the advantages of a high reduction ratio is that it enables a fine adjustment of the top dead centre (TDC) of the main pistons7. Another advantage of a high reduction ratio is that it eliminates the possibility of the output (worm gear13,14) driving the input (worm control shaft11,12), also known as a self-locking configuration.

Split Cycle Process

The inventive use of the know split cycle process in the present invention comprises a two-stage compression and a two-stage expansion. Said stages are split between main cylinders I, III; II, IV and auxiliary cylinders V, VII; VI, VIII. In the embodiment disclosed in the figures, the engine has four main cylinders I, III; II, IV and four auxiliary cylinders V, VII; VI, VIII. As an alternative embodiment, it would be possible to double the number of cylinders by adding them in series or in parallel.

FIG. 6shows a vertical section view of the engine, showing the complete right engine side R, and the left engine side L with most of the static parts hidden, leaving the valve arrangement, pistons and auxiliary cylinder liners67. The section view cuts through the centre of the auxiliary yoke3and the four auxiliary cylinders V, VII; VI, VIII.

Within each auxiliary cylinder V, VII; VI, VIII, the auxiliary piston8defines an outer space and an inner space, wherein the inner space, closest to the auxiliary yoke3, is used for compression and the outer space is used for expansion. The pressure difference between the outer space and the inner space of the auxiliary cylinder V, VII; VI, VIII is up to approximately 6 bar at full power. The auxiliary pistons8are made of a material (preferably steel) with mechanical and thermal properties allowing some hot gas leakage from the outer space to the inner space without causing erosion of the auxiliary pistons8. The auxiliary pistons8are therefore equipped with a number of pressure trap grooves72instead of piston rings. The clearance between the auxiliary piston8and the auxiliary cylinder liner67is very small. The centring of the pistons8is secured as their auxiliary piston rods4are centric stable. Fluids slipping inn between the auxiliary piston8and the auxiliary cylinder liner67will be trapped in the pressure trap grooves72. It is also acceptable if some fluids travel from one side of the auxiliary piston8to the other. This design eliminates mechanical friction loss in the auxiliary cylinders8, and they do not require lubrication.

Two auxiliary cylinders V, VII; VI, VIII of the same engine side R, L are equipped with a pair of oppositely directed check valves69,70. Fluids can flow into the inner space through a first check valve69arranged in a first auxiliary cylinder V, VII; VI, VIII. As vacuum builds up in the inner space, the first check valve69will open and allow fluids to enter. The first check valve69is an inlet into the inner space, which prevents fluids from escaping the inner space. Through a second check valve70arranged in a second auxiliary cylinder V, VII; VI, VIII, fluids can escape the inner space. As pressure builds up in the inner space, the second check valve70will open and allow fluids to escape. The second check valve70is an outlet from the inner space, which prevents fluids from entering the inner space. Fluid communication is provided between the inner spaces of the first and second auxiliary cylinders V, VII; VI, VIII by an interconnecting bore105, casing or similar (also illustrated inFIG. 7). The check valves69,70are positioned at the bottom of each auxiliary cylinder V, VII; VI, VIII, which is the end closest to the yoke3. In the centre of the check valves69,70, an aperture is provided having a sealing interface towards the reciprocating auxiliary piston rods4. The check valves69,70can for instance comprise discs sealing the bottom of the auxiliary cylinders V, VII; VI, VIII, which discs are spring-loaded in the desired direction to a suitable preload.

This design makes the combined inner spaces of an auxiliary cylinder pair V, VII; VI, VIII of the same engine side R, L substantially sealed, which in turn enables suction of ambient air into the inner space by the auxiliary pistons8, and it also enables compression of said ambient air by said auxiliary pistons8. The flow of ambient air into the inner space is regulated by a throttle63. Compressed air/fuel mixture escaping the inner space of the auxiliary cylinders V, VII; VI, VIII through the second check valve70is led through a connection channel62into an inlet manifold of the main cylinders I, III; II, IV of the opposite engine side R, L. The charge of compressed air/fuel mixture will enter a first main cylinder I, III; II, IV having an open suction valve31, a second main cylinder I, III; II, IV will at this point have a closed suction valve31. At full throttle, the filling ratio in a main cylinder I, III; II, IV will be up to 200%. The main cylinder I, III; II, IV receiving the charge will be at its BDC. Once the charge is received in the main cylinder I, III; II, IV, the suction valve31will close and the main piston7will compress the charge further within said main cylinder I, III; II, IV; hence, a two-stage compression. The consecutive charge delivered to said inlet manifold will be received by a second main cylinder I, III; II, IV, this time with an open suction valve31, and the first main cylinder I, III; II, IV having a closed suction valve31.

The main scotch yokes110are arranged synchronized on the crankshaft1and the auxiliary scotch yoke120is arranged 180° offset on the crankshaft1. This means that when the main pistons7of an engine side R, L is at the TDC, the auxiliary pistons8of the same engine side R, L is at the BDC. Table 1 shows the steps taking place in all cylinders I, III; II, IV, V, VII; VI, VIII during a complete cycle.

FIG. 7shows a 90° section cut of a top section of the engine. The figure illustrates a cylinder bottom plate52, a cylinder block81, a valve seat plate54, a metal gasket55and a valve top block56, where the section cut goes through the centre of both a main cylinder I, III; II, IV and an auxiliary cylinder V, VII; VI, VIII, both with their pistons7,8and piston rods4,5removed.

After the second stage of the two-stage compression has been completed in a main cylinder I, III; II, IV, the charge is ignited by a spark plug47. An expansion then takes place in the main cylinder I, III; II, IV, like in an ordinary internal combustion engine. When the expansion has driven the main piston7to its BDC, there will remain some pressure in the exhaust gas inside the main cylinder I, III; II, IV. This remaining pressure is then transferred to the auxiliary cylinders V, VII; VI, VIII for a second expansion stage; hence a two-stage expansion. Said expansion takes place in a combined outer space of an auxiliary cylinder pair V, VII; VI, VIII of the same engine side R, L, driving the auxiliary pistons8from their TDC to their BDC.

Between the cylinder block81and the valve top block56, a valve seat plate54is arranged. This valve seat plate54enables the fluid transfer from the main cylinders I, III; II, IV to the auxiliary cylinders V, VII; VI, VIII of the same engine side R, L.FIG. 8a and b shows both sides of the valve seat plate54. A valve seat plate54is provided on each engine side R, L. Each valve seat plate54interface two main cylinders and two auxiliary cylinders V, VII; VI, VIII. For the main cylinders I, III; II, IV, the valve seat plate54provides a suction valve seat101, a discharge valve seat102and a spark plug seat104. For the auxiliary cylinders V, VII; VI, VIII, the valve seat plate54provides a fluid transfer channel100aand an exhaust valve seat103. Said fluid transfer channel100ais interconnecting both auxiliary cylinders V, VII; VI, VIII with each other and with both main cylinders I, III; II, IV of the same engine side R, L. The fluid transfer channel100ais a groove machined into the backside of the valve seat plate54, sealed off by a metal gasket55. The communication between the transfer channel100aand the main cylinders I, III; II, IV are controlled by the discharge valves32, whilst the communication between the transfer channel100aand the auxiliary cylinders V, VII; VI, VIII are permanently open through a transfer inlet (100b).

Once the first expansion stage is completed in a first main cylinder I, III; II, IV, its discharge valve32opens. At this point, the main piston7of said main cylinder is at its BDC, and the auxiliary pistons8of the same engine side R, L are at their TDC. Exhaust gas is transferred from the main cylinder I, III; II, IV to the auxiliary cylinders V, VII; VI, VIII via the transfer channel100a. The second expansion stage takes place inside the outer space of the auxiliary cylinders V, VII; VI, VIII. The second expansion stage is completed when the auxiliary pistons8reach their BDC. At that point, the discharge valve32of the main cylinder I, III; II, IV closes, and the exhaust valves33of the auxiliary cylinders V, VII; VI, VIII open. Exhaust gas escapes through the exhaust valves33of the auxiliary cylinders V, VII; VI, VIII, into the exhaust manifold65. A first part of said exhaust manifold65being included in the valve top block56. When the auxiliary pistons8reach their TDC again, all exhaust has escaped the auxiliary cylinders V, VII; VI, VIII and the exhaust valves33close. The auxiliary cylinders V, VII; VI, VIII will then receive a new pressurized exhaust gas from a second main cylinder I, III; II, IV of the same engine side R, L. The second expansion stage drives the first compression stage and powers the crankshaft1.

The cylinder bottom plate52has apertures for the main piston rods5and the auxiliary piston rods4to pass through. In the areas of the cylinder bottom plate52interfacing the main cylinders I, III; II, IV, additional apertures are provided for the passage of air.

Variable Valve Timing

FIGS. 9 and 10illustrate the mechanism enabling variable valve timing in the present invention. Rotational movement of the crankshaft1is transferred to the two camshafts30by means of interconnected gears16,17a,17b,41and connection shafts44,45. By longitudinally adjusting a connection shaft44,45, the rotation of the corresponding camshaft30will be altered relative to the rotation of the crankshaft1, i.e. the timing of the opening/closing of valves will change relative to the travel of the corresponding pistons.

FIG. 9shows a horizontal section view of the right engine side R with all components present, and a top view of the left engine side L with most static components removed. The section view cuts through the centre of the main cylinders I, III and the centre of the connection shaft44.

FIG. 10shows an isometric view of the engine with the right engine side R having most static components removed, and a substantially complete left engine side L.

The gear ratio between the crankshaft1and the camshafts30is 2:1, i.e. the camshaft30will turn one revolution as the crankshaft1turns two revolutions. During two revolutions of the crankshaft1, the main cylinders I, III; II, IV will performs a complete cycle (four strokes). The auxiliary cylinders V, VII; VI, VIII will perform a complete cycle as the crankshaft1turns one revolution. Because the suction valves31, discharge valves32and exhaust valves33of the same engine side R, L are operated by the same camshaft30, a 180° double cam74, driving the exhaust valve33, is positioned in the middle part of the camshaft30.

In a first end of the crankshaft1a flywheel61is arranged, in a second end of the crankshaft1a crankshaft bevel gear16is arranged. In one end of the camshafts30, oriented in the same direction as the second end of the crankshaft1, a camshaft bevel gear41is arranged. A first connecting shaft bevel gear17ain engagement with the crankshaft bevel gear16, arranged in a 90° configuration, lines up with a second connecting shaft bevel gear17bin engagement with the camshaft bevel gear41, arranged in a 90° configuration. Said connecting shaft bevel gears17a,17beach have a centrally protruding, relatively short, spindle42a42bwith external helical splines20a,20b. A first spindle42ahaving left hand external helical splines20a, and a second spindle42bhaving right hand external helical splines20b, or vice versa. Said spindles42a,42bare concentrically oriented and directed towards one another. A connection shaft44,45connects the two connecting shaft bevel gears17a,17bof the same engine side R, L. The connecting shaft44,45has internal helical splines22a,22bcorresponding to those on the spindles42a,42b. Where a first end of the connection shaft44,45has right hand internal helical splines22a, and a second end of the connection shaft44,45has left hand internal helical splines22b, or vice versa. Lengthwise the connection shaft44,45is shorter than the distance between the two connecting shaft bevel gears17a17b. The length of the connection shaft44,45is long enough to always be engaged with both spindles42a42b, but short enough to allow some play in its longitudinal direction.

For simultaneous axial movement of the two connection shafts44,45, they are longitudinally interconnected. Adjustment of the connection shafts44,45may be operated by hydraulic or electric linear actuators.

LIST OF REFERENCE NUMERALS