TurbochargerTwo stroke crosshead engines must be supplied with air above atmospheric pressure for it to work. Although turbochargers were developed in 1925, it was not until 1950s that large two stroke engines were turbocharged. Pressurized air is needed to scavenge the cylinders of the exhaust gases and supply the charge of air for next combustion cycle was provided by mechanically driven air compressors (roots blower) or by using the space under the piston as a reciprocating compressor (under piston scavenging). This of course meant that the engine was supplying the energy to compress the air, which meant that the useful work obtained from the engine was reduced by this amount.

About 35% of the total fuel energy goes out in the exhaust gas. The turbocharger uses 7% of the total energy (20% of the exhaust gas energy) to drive a single row turbine. The turbine shaft drives a rotary compressor. Air is drawn and compressed. Due to compression, the air temperature rises. Hence it is cooled in a cooler to increase its density and then sent to the air inlet manifold or scavenge air receiver. At full power of diesel engine, the turbocharger may be rotating at > 10000rpm.

Power of a two stroke diesel engine = pm x L x A x N x no. of cylinders Where pm = mean effective pressure; L = stroke of the engine;A = cross sectional area of the cylinder; N = revolution per second of the engineThus to increase the power of the engine of given swept volume i.e. the power to weight and volume ratios of the engine we have to increase either mean effective pressure or the revolution per second of the engine. The approach of increasing power output by increasing speed is unattractive, due to rapid rise of mechanical and aerodynamic losses, and the corresponding fall in brake thermal efficiency.

For increasing the mean effective pressure, more fuel has to be burnt during the cycle of the engine, which requires higher quantity of air per cycle of the engine. The purpose of supercharging is to increase the mass of air trapped in the cylinders of the engine, by raising its density. A compressor is used to achieve the increase in air density. Two methods of supercharging can be distinguished by the method to drive the compressor. If the compressor is driven from the crankshaft of the engine, the system is called mechanically driven supercharging or often just supercharging. If a turbine drives the compressor, which itself is driven by the exhaust gas from the cylinders, the system is called turbocharging. The shaft of the turbocharger links the compressor and the turbine, but is not connected to the crankshaft of the engine.

The advantage of the turbocharger, over a mechanically driven supercharger, is that the power required to drive the compressor is extracted from the exhaust gas energy rather than the crankshaft. Thus turbocharging is more efficient than mechanical supercharging. However the turbine imposes a flow restriction in the exhaust system, and therefore the exhaust manifold pressure will be greater than atmospheric pressure. If sufficient energy can be extracted from the exhaust gas, and converted into compressor work, than the system can be designed such that the compressor delivery pressure exceeds that at turbine inlet, and the inlet and exhaust processes are not adversely affected. For a compressor pressure ratio of 5, allows to increase the specific power output of the engine by 400%.

Turbocharging systems principles

Consider now an ideal turbocharged four-stroke engine. Turbocharging raises the inlet manifold pressure, hence the inlet process (12-1) is at pressure P1, where P1, is above ambient pressure Pa. The blow-down energy is represented by area 5-8-9. The exhaust manifold pressure (P7) is also above the ambient pressure Pa. The exhaust process from the cylinder is represented by line 5,13,11, where 5,13 is the blow-down period when the exhaust valve opens and high pressure gas expands out into the exhaust manifold.

Process 13, 11 represents the remainder of the exhaust process, when the piston moves from BDC to TDC displacing most of the gas from the cylinder to exhaust manifold. This gas is above ambient pressure and therefore also has the potential to expand down to ambient pressure whilst doing useful work. The potential work that could be done is represented by the hatched area 13-9-10-11. This work is done by the piston but could be recovered by the turbine in the exhaust. It will be called the piston pumping component of the exhaust energy.The maximum available energy to drive a turbine will be sum of areas 5-8-9 and 13-9-10-11, but it is impossible to devise a practical system that will harness all this energy. To achieve this, the turbine inlet pressure must instantaneously rise to P5 when the exhaust valve opens, followed by isentropic expansion of the exhaust gas through P7 to the ambient pressure (P8 = Pa). During the displacement part of the exhaust process, the turbine inlet pressure would have to be held at P7. Such a series of process is impractical.

Consider a simpler process that would occur if a larger chamber were fitted between the engine and the turbine inlet in order to damp down the pulsation in exhaust gas flow. The turbine acts as a flow restrictor creating a constant pressure (P7) in the exhaust manifold chamber. The available energy at the turbine is given by area 7-8-10-11. This is the ideal constant pressure turbocharging system. Next consider an alternative system, in which a turbine wheel is placed directly downstream of the engine, very close to the exhaust valve. The gas would expand directly through the turbine along line 5-6-7-8, assuming isentropic expansion and no losses in the exhaust port. If the turbine were sufficiently large, both cylinder and turbine inlet pressure would drop to equal ambient pressure before the piston has moved significantly from BDC. Thus the piston pumping work would be zero during the ideal exhaust stroke and area 5-8-9 would represent the available energy at the turbine. This is the ideal pulse turbocharging system.

Principles of constant pressure turbocharging With constant pressure turbocharging, the exhaust ports from all cylinders are connected to a single exhaust manifold, whose volume is sufficiently large to ensure that its pressure is virtually constant. The unsteady exhaust flow processes at the cylinders are damped into a steady flow at the turbine. Only one turbocharger need be used, with a single entry from the exhaust manifold, but frequently several smaller units are fitted so that a reasonable boost pressure can be obtained in the event of a turbocharger failure. A major advantage of the constant pressure system is that turbine inlet conditions are steady and known, hence the turbine can be matched to operate at optimum efficiency at specified engine conditions. The main disadvantage is that the available energy entering the turbine is low, since full advantage has not been taken of the pulse energy. In Figure4 area 7-8-10-11 denotes energy available to the turbine, hence the energy represented by area 5-7-13 cannot be used. This energy is not lost, since energy loss only occurs by heat transfer, but since no work is done during the pressure reducing process 5-7 it represents a loss of potential turbine work.

Typically, a constant pressure exhaust manifold will consist of a large diameter pipe running along the exhaust side of an engine, with each exhaust port connected to it via a short stub pipe. On a vee engine, the large bore manifold will usually lie between the banks with the inlet valves and manifolds arranged to be on outside.The volume of the exhaust manifold should be sufficient to damp pressure pulsations down to a low level. For guidance however, it can be stated that the volume would normally be in the range of 1.4 to 6 times the total swept volume of the engine.If the exhaust manifold volume is not sufficiently large, the blow-down or first part of the exhaust pulse from a cylinder will raise the general pressure in the manifold. Since all cylinders are connected to the same manifold, it is inevitable (if the engine has more than three cylinders) that at the moment when the blow-down pulse from one cylinder arrives in the manifold, another cylinder is nearing the end of its exhaust process. The pressure in the latter cylinder will be low; hence any increase in exhaust manifold pressure will impede its exhaust process. This will be particularly important where the cylinder has both inlet and exhaust valves or ports partially open and are relying on a through-flow of air for scavenging the burnt combustion products. A rise in exhaust manifold pressure at this time is virtually inevitable in an engine with more than three-cylinders, unless the volume is large. This will be particularly important on a two-stroke engine, since if the exhaust pressure exceeds inlet pressure during scavenging, the engine cannot run at all.

From a purely practical point of view, the exhaust manifold is simple to construct although it may be rather bulky, particularly relative to small engines with few cylinders. However, for large engines with many cylinders, the convenience of being able to join all cylinders to a common exhaust manifold with a single turbocharger on top or at either end is useful. A major disadvantage of the constant pressure system arises from the use of an exhaust manifold having large volume. When the engine load is suddenly increased or a rapid engine speed increase is required, the pressure in the large volume is slow to rise. Hence the energy available at the turbine increases only gradually. Turbocharger, and therefore engine response, will be poor. The poor response of the constant pressure turbocharging system restricts it from consideration for applications where frequent load (or speed) changes are required.The turbine inlet temperature (T3), the inlet manifold pressure (P2) and exhaust manifold pressure (P3) are functions of overall turbine efficiency () and to lesser extent, the air-fuel ratio (AFR) also plays its part. The air-fuel ratio at full load will be governed by thermal loading or onset of black smoke in the exhaust. The turbine inlet temperature will also be dependent on air-fuel ratio, the amount of cool scavenge air passing through the cylinders and heat loss from the exhaust manifold.

Principles of pulse turbocharging

Here the effect of varying pipe length on the timing of the reflected pulse is shown. In case 1, the reflection occurs after the exhaust valve has closed causing no problem, but this is a rare case since it can occur with an exceptionally long manifold. More common is case 4, in which the reflection time is very short relative to the valve opening period. Case 2 is the serious one, that can occur with long pipes, of the reflected pulse raising exhaust pressure at the valve or port, during the scavenge period. The turbocharger position and exhaust pipe length must be chosen to avoid this situation, or scavenging will be seriously impaired.

Principles of pulse turbocharging

On a multi-cylinder engine, narrow pipes from several cylinders can be connected via a single branch manifold to one turbine. Consider a three cylinder four-stroke engine. Due to the phase angle between cylinders the opening periods of the exhaust valves follow successively every 2400 with very little overlap between them.

Thus a steady train of pressure pulses arrives at the turbine, virtually eliminating the long periods of pure windage, although the average turbine efficiency will remain lower than that obtained with correctly matched constant pressure system (operating near the peak of the efficiency curve). The remaining important point to consider is the exhaust pressure close to the valve, during the valve overlap (scavenging) period.

As with the constant pressure system, a good pressure drop between inlet and exhaust manifold during the period when both valves are open is important in the case of a four-stroke engine with significant valve overlap and vital for a two-stroke engine. The pressure history in the inlet and exhaust manifolds (at a valve) is shown, the pressure drop during the period of valve overlap being hatched. Clearly, at the running condition shown, the pressure drop is satisfactory.The diagrams have shown the way to increase average turbine efficiency by reducing windage periods, whilst avoiding interference with scavenging of one cylinder due the effect of the blow-down pulse from another. The pressure pulse exhausting from a cylinder travels along the manifold until it reaches a junction. At the junction it divides into two pulses (each of smaller magnitude due to the effective area increase) one travelling down each adjacent pipe. One pulse will travel towards the turbine; the other will arrive at the exhaust valve of another cylinder. It is the latter pulse, from cylinder number 3, that has arrived near cylinder 1 just at the end of the scavenge period of cylinder 1, that could be a problem. If it had arrived earlier (perhaps due to shorter exhaust pipes) it would have interfered with scavenging. This type of interference due to the direct action of a pressure wave from another cylinder is quite separate from the action of a pressure pulse reflected from the turbine, whether the latter started from cylinder 1,2 or 3.

Most engines have four or more cylinders. Let us consider a six cylinder engine. If all six cylinders were connected to a single entry turbine via narrow pipes, the pressure waves from each cylinder would significantly interfere with the exhaust processes of each other during valve overlap and the exhaust stroke, thus increasing piston pumping work. The effect would be poor engine efficiency. A two-stroke engine might not operate under these conditions.

The difficulty can be avoided by simply connecting the cylinders in two groups of three, either to two different turbines, or separate entries of a single turbine. If the correct cylinders are grouped together, then the pressure pulse system in each group will be the same as that shown in earlier figure. It may be concluded that the six-cylinder engine is similar to the three-cylinder, from the turbocharging point of view, but turbine performance may be slightly worse due to the losses associated with the join of two sectors of a divided entry turbine.

It is disadvantageous to connect more than three cylinders to a single turbine entry. Thus for the four-cylinder engine, pairs of cylinders (1-2 and 3-4) would be connected to a double entry turbine. On engines with other numbers of cylinders, the general rule will be to connect cylinders whose firing sequences are separated by 2400 crank angle (in case of four-stroke) and 1200 (two-stroke) to a turbine inlet, and select those cylinders whose exhaust processes are evenly spaced out. However, this is not always possible. For example, on a vee-form engine, the vee angle will introduce an additional phase difference to the firing intervals between cylinders. The principal advantage of the pulse over the constant pressure system is that the energy available for conversion to useful work in the turbine is greater. However, this is of little value if the energy conversion process is inefficient.

With three cylinders to a turbine entry the average turbine efficiency will be much higher since windage is almost eliminated. The efficiency is better still if the valve timing permits a larger overlap by having longer exhaust periods (2900) as is the practice in medium speed diesel engines. However, turbine efficiency, averaged over the unsteady flow cycle, will be lower than obtained in a well matched steady flow system. If two cylinders were connected to a turbine entry the average turbine efficiency will be lower than would be the case with three cylinders, since (short) windage periods would exist . Thus pulse turbocharging system is most suitable for those engines whose exhaust manifolds may connect groups of three cylinders to a turbine entry. However, even if this is not possible, the loss in turbine efficiency due to partial admission and unsteady flow is usually more than offset by the additional energy available at the turbine, hence the pulse system is by far the more widely used. In practice, the constant pressure system is used exclusively on very large, highly rated two-stroke engines. On these engines the ratings are such that very large pressure pulses would be generated with the pulse system. Since most of the exhaust pulse energy coincides with the peak of the pulse, matching this point with high instantaneous turbine efficiency is important. In practice it is difficult to maintain high turbine efficiency when the pressure ratio exceeds 3:1, hence turbine efficiency will be low if exhaust pressure pulse amplitude substantially exceeds this value. This is what happens on very highly rated engines; hence constant pressure systems operate with higher turbine efficiency, more than offsetting their lower available energy.

Pulse turbocharging of two-stroke engines The major difference of 2-stroke engine with respect to 4-stroke engine is that the exhaust pressure diagram is somewhat different, due to the long scavenge period of the two-stroke engine. It tends to consist of quite distinct blow-down and scavenge period.

Pulse turbocharging

Constant pressure turbocharging

AdvantagesDisadvantages1. High available energy at turbine1. Poor turbine efficiency with one or two cylinders per turbine entry2. Good performance at low speed and load2. Poor turbine efficiency at high ratings3. Good turbocharger acceleration3. Complex exhaust manifold with large number of cylinders4. Possible pressure wave reflection problems on some engines

AdvantagesDisadvantages1. High turbine efficiency due to steady flow1. Low available energy at turbine.2. Good performance at high load2. Poor performance at low speed and load 3. Simple exhaust manifold3. Poor turbocharger acceleration

Compressor characteristic and the surge limit

Centrifugal compressor characteristics are similar to those of centrifugal pumps. At a constant RPM, the characteristic would appear similar to the figure. At constant speed the discharge pressure first rises as volumetric flow increases and then drops off rather sharply. The compressor efficiency curve also rises to a peak, although at any constant this peak is to the right of the pressure peak. The power consumed by the compressor is related to the product of discharge pressure and flow rate.

In the region to the right of the peak in pressure curve, operation will be stable: in this region a momentary drop in volumetric flow rate, for example, perhaps brought on by a momentary reduction in engine speed, will be countered by a rise in pressure, with little or no effect on the turbine. In the region to the left of the pressure peak, a momentary drop in volumetric flow rate will be accompanied by a drop in discharge pressure and a reduction in compressor power consumption. Operation in the unstable area to the left of the pressure peak may result in compressor surge. As the pressure at the compressor discharge falls below that downstream, the flow can reverse. The result can simply be a pulsation if the situation is not severe or of long duration, or the reversed flow can continue to the air intake and become audible, ranging in volume from a soft sneezing to a very loud backfiring sound.

Obviously, operation in the surge region should be avoided; consequently, turbocharger designers establish a line, called a surge limit, through the pressure characteristics slightly to the right of the peak. Similar data as previous figure are obtained at several constant speeds covering the range of operation, and plotted together on the same axes. The resulting compressor performance map is shown.

Turbocharger matchingInlet conditions of the compressor P1 (after pressure drop across air suction filter) and T1 are selected.An estimate is made of the power of the engine at a particular engine RPM. Also, an estimate is made of amount of air, ma, the engine would require at above conditions. ma = Vsw x a x vol x N (for two stroke diesel engine). The pressure in the air manifold and temperature are estimated to get the a.

A drop across air cooler is assumed and added to air manifold pressure to get the compressor discharge pressure P2. The compressor pressure ratio P2/P1 can now be calculated.The compressor frame size and its diffuser can now be selected by entering the family of compressor maps (Figure 20 and Figure21) with the values of P2/P1 and Va or (ma T0/P0). The operating point must have adequate margin from surge limit i.e. it must be 10% to 20% to the right of the surge limit at the value of P2/P1.

From the performance map of the selected compressor and diffuser, the compressor efficiency (c) and turbocharger RPM are read at the operating point.The required power to drive the compressor is given by Wc = ma Cpa T1 [ (P2/P1) (-1)/ -1]/ c The required turbine power Wt = Wc / mech where mech is the known characteristic of the bearings.Again an estimate is made of the gas conditions at the turbine inlet i.e. P3 and T3 from basic principles and empirical data including previous performance.The turbine outlet pressure P4 is also estimated by adding an amount to atmospheric pressure to allow for typical losses through exhaust gas economiser. Then expansion ratio P3/P4 is calculated.The mass flow rate of gas, mg, is calculated by adding mass flow rate of fuel, mf, to mass flow rate of air, ma, which was estimated previously.

In general, the selection of a compressor wheel diameter predetermines turbine characteristics, which includes wheel diameter and blade length. With values of P3/P4 and Vg, a turbine blade and nozzle angle selection curve, such as Figure 22, can be entered for the frame size under consideration, to select nozzle opening and blade angle.

The following are calculated: Mean tangential velocity of blade, Um = x Dm x RPM where Dm = mean diameter of turbine wheel Ideal gas speed at nozzle exit Cg = [2Cpg T3 {1- (P4/P3)-1/]The turbine efficiency t can then be obtained from figure 23.The available power of the turbine Wt = m g C pg T3 [ 1 (P4 / P3 ) (-1)/ ] t

A comparison is made between the turbine power available and turbine power required. If available power is greater than required power, then the estimate of air manifold pressure, made in step 2, can be raised and the procedure is repeated with this new assumption. If the available power is lower, a lower value of air manifold pressure is assumed.This iterative process is continued till the required turbine output matches the achieved value within a percent or two, with the final matching to be done by actually running the turbocharger and engine on a test-bed and making final adjustments by changing compressor diffuser vanes and turbine nozzle rings.

Cleaning Turbochargers in operationPeriodic cleaning reduces or even prevents contamination, allowing significantly longer intervals between overhauls. The proposed cleaning method, carried out periodically, will prevent a thick layer of dirt from forming. A thick layer of dirt can cause a drop in efficiency and increased unbalance on the compressor side of the turbocharger, which could influence the lifetime of the bearings.The compressor wheel of the turbocharger can be cleaned during operation by spraying water into the air inlet casing. The dirt layer is removed by the impact of the injected water. Since the liquid does not act as a solvent there is no need to add chemicals. The use of saltwater is not allowed, as this would cause corrosion of the aluminium compressor wheel and the engine. Water is injected from a water vessel that holds the required quantity of water.

ProcedureThe best results are obtained by injecting water during full-load operation of the engine, i.e. when the turbocharger is running at full speed. The complete contents of the water vessel should be injected within 4 to 10 seconds. Successful cleaning is indicated by a change in the charge air or scavenging pressure, and in most cases by a drop in the exhaust gas temperature. If cleaning has not produced the desired results, it can be repeated after 1 0 minutes. The interval between compressor cleanings will depend on the condition of the turbocharger suction air. It can vary from 1 to 3 days of operation. If a very thick layer has built up and it cannot be removed using the method described, it will be necessary to dismantle the turbocharger in order to clean the compressor side. Since the dirt layer is removed by the kinetic energy of the water droplets, the engine has to be run at full load.

Cleaning the TurbineThe combustion of heavy fuel in diesel engines causes fouling of the turbine blades and nozzle ring. The result of this fouling is reduced turbine efficiency and engine performance as well as an increase in the exhaust gas temperature, Experience has shown that the contamination on the turbine side can be reduced by regular cleaning in operation, and that such cleaning allows longer intervals between turbocharger overhauls. Some of the deposits have their origin in soot, molten ash, scale and unburned oil, partially burnt fuel and sodium vanadyl-vanadat. Investigations have shown that most of the residues are caused by the calcium in the lube oil reacting with the sulfur from the fuel to form calcium sulfate during the combustion process.

The quantity of the deposits depends on the quality of the combustion, the fuel used, and the lube oil consumption. The frequency with which cleaning has to be carried out depends on the extent of the contamination on the turbine side. Two cleaning methods exist:Wet cleaning (water injection)dry cleaning (solid particle injection)

Procedure for wet cleaningThe exhaust gas temperature before the turbine should be in the range of 200 to 4300 CThe boost pressure should be above 0.5 bar to prevent water entering the oil chamber on the turbine side.The quantity of injected water will depend on the exhaust gas temperature, water pressure, size of the turbo-charger and number of gas inlets.

Water should be injected for 5 to 10 minutes. Check if the water has entered the turbine parts by opening the drain of the gas outlet casing. Water flowing out provides assurance that enough water has passed the nozzle ring and the turbine blades. The interval between turbine cleanings will depend on the combustion, the fuel used and the fuel oil consumption. It can vary from 1 to 20 days of operation. PrincipleThe dirt layer on the turbine components is removed by thermal shock rather than the kinetic energy exerted by the water droplets.

Procedure for dry cleaningThe exhaust gas temperature before the turbine should not exceed 5800 C.Dry cleaning has to be carried out more often than water cleaning as it is only possible to remove thin layers of deposits. A cleaning interval of 1 to 2 days is recommended.To ensure effective mechanical cleaning, granulated dry cleaning media are best injected into the turbine at a high turbocharger speed.The quantity needed will vary from 0.2 l to 3 l, depending on the size of the turbocharger.

Experience has shown that the best results are achieved with crushed nut-shell or granulate.PrincipleThe layer of deposits on the turbine components is removed by the kinetic energy of the granulate causing it to act as an abrasive. Experience has shown a combination of the two to be very effective, especially in the case of 2-stroke engines.