GB2263308A

GB2263308A - Differential drive for supercharged engine

Info

Publication number: GB2263308A
Application number: GB9200911A
Authority: GB
Inventors: Ian Charles Crossley
Original assignee: Individual
Current assignee: Individual
Priority date: 1992-01-16
Filing date: 1992-01-16
Publication date: 1993-07-21
Also published as: GB9200911D0

Abstract

A supercharging compressor 6, for an i.c. engine, is driven via a planetary gear train, the planet gear carrier 1, of which is driven by the engine, and the ring gear 3, of which is driven by an exhaust driven turbine 4, via pinion 2. In an alternative arrangement (figure 2) speed increasing gears (8, 9) are interposed between the turbine and the pinion 2 and between the sun gear 5 and the compressor 6. At low engine speeds the compressor is able to be driven only by the engine at its maximum speed and hence its maximum boost pressure. The turbine can be used to maintain the compressor at maximum speed and maximum boost pressure, throughout the useful engine operating speed range. The turbine also provides part of the power to drive the compressor, and augments the power produced by the engine. It is stated that the turbine could be replaced by an electric, hydraulic or pneumatic motor. <IMAGE>

Description

DIFFERENTIAL TURBOCHARGER This invention relates to the supercharging, or charge air compression, of internal combustion engines.

Charge air compression in an internal combustion engine allows the power for a given engine capacity, or swept volume, to be substantially increased.

Displacement compressors of the sliding vane type or rotary lobe type have been much used in the past, but tend to be large and heavy, and of lower efficiency than centrifugal compressors for the same application.

Centrifugal charge air compressors are small and light, and are efficient and effective at full speed. However, when directly geared to an engine, they produce very little boost pressure at low and medium engine speeds, because their operating characteristics are not well matched to the requirements of fixed displacement internal combustion engines.

Turbochargers, in which a centrifugal charge air compressor is driven directly by a centripetal turbine extracting power from the engine exhaust gas, are currently favoured. Turbochargers are not mechanically linked to the engine. This fact allows the compressor to run closer to optimum speed, and thus produce a charge air boost pressure, which is more closely matched to engine requirements.

However, because of the absence of any mechanical link, turbochargers tend to be slow in responding to engine throttle demand. "Turbo lag", as it is known, can be of significant disadvantage. By careful design of the engine system and its components, turbo lag can be minimised but not completely eliminated. Design compromises necessary to minimise turbo lag may adversely affect other aspects of engine performance, such as thermal efficiency, or specific power output.

According to the present invention there is provided a differential drive system for centrifugal charge air compressors for internal combustion engines, in which the compressor speed, and therefore the boost pressure developed by the compressor, can be controlled throughout the useful engine speed range. The system uses an exhaust gas turbine to modulate the effective speed ratio between the engine and the compressor, to give control of the compressor speed.

Modulation of the compressor speed enables the engine torque to be optimised throughout the engine speed range. This offers the possibility of an extremely flexible engine for applications requiring variable speed operation, such as, for example, road or rail transport. The improvement in flexibility might well reduce the complexity and reduce the number of gears required in the main transmission gear box.

The differential system proposed for speed modulation of the compressor offers another benefit, in that surplus power from the exhaust gas turbine which is not required to drive the compressor, augments the engine power. The system can therefore improve the specific fuel consumption, and also the specific power output of the engine.

A specific embodiment of the invention will now be described with reference to Figure 1. Perspective view Figure 2. Gear train arrangement with additional speed increasing gears.

Figure 3. Example of compressor speed modulation by turbine.

Figure 4. Further examples of compressor speed modulation by turbine.

Figure 1. shows a perspective view of a differential drive system in which the differential effect is achieved with an epicyclic gear.

In Figure 1. the engine drives the planet gear carrier 1 of the epicyclic gear system. The compressor 6 is driven through the sun wheel 5 of the epicyclic gear system.

At low engine speeds the ring gear 3 of the epicyclic is required to be kept stationary, or only driven at very low speed, so that the compressor is driven primarily by the engine.

By selecting appropriate gear sizes, the sun wheel, or compressor pinion, can achieve its maximum operating speed at a relatively low engine speed.

If the ring gear of the epicyclic gear system is then arranged to rotate at some constant speed, this will change the effective gear ratio between the engine and the compressor.

To facilitate this effect, the ring gear of the epicyclic system is equipped with internal and external teeth, and according to the present invention, an exhaust gas turbine 4 is arranged to drive the ring gear through a pinion 2 meshing with the external teeth. For a system constructed as in Figure 1.

rotation of the turbine by the engine exhaust gas will have the effect of slowing down the compressor.

In this way, the effective gear ratio between the engine and the compressor can be modulated in accordance with the ring gear speed, and consequently therefore, in accordance with the turbine speed.

By the above means the boost pressure of the centrifugal charge air compressor, and consequently also the engine torque, can be maximised over the whole of the useful operating speed range of the engine.

Maximum boost pressure for a given engine speed, is defined as the highest boost pressure which the compressor can achieve, at the charge air suction volume flow dictated by the engine speed, whilst avoiding the compressor surge line.

The shaft speeds of the differential gear system shown in Figure 1. are given by the equation : nc(Dc) = ne(Dp + Dc) - nt(Dt)Dri/Dro Where nc = compressor pinion speed nt = turbine pinion speed ne = planet carrier speed Dc = compressor pinion pitch circle diameter Dt = turbine pinion pitch circle diameter Dp = planet wheel pitch circle diameter Dri = inner ring gear pitch circle diameter Dro = outer ring gear pitch circle diameter For vehicle applications there may be a requirement for additional speed increasing gears between the compressor pinion of the differential gear system, and the compressor itself, and also between the turbine pinion of the differential gear system and the turbine itself, in order that the compressor and the turbine can run at suitably high speeds.

Similarly, it may be convenient to drive the planet carrier of the differential gear, through a speed increasing system to assist in achieving this objective.

In this latter case, the speed increasing system may also allow the differential turbocharger unit to be located away from the engine crankshaft, if the drive is via chains or flexible belts for example.

Where additional speed increasing systems are employed,the following equations will apply Nc = (rc)nc Nt = (rt)nt Ne = ne/re Where Nc = compressor speed Nt = turbine speed Ne = engine speed rc = speed ratio compressor to compressor pinion rt = speed ratio turbine to turbine pinion re = speed ratio engine to planet carrier The relationship between engine, compressor, and turbine speeds for specific epicyclic gear dimensions is calculated in the following example :: compressor pinion pitch circle diameter Dc = 30 mm turbine pinion pitch circle diameter Dt = 30 mm planet wheel pitch circle diameter Dp = 45 mm inner ring gear pitch circle diameter Dri = 120 mm outer ring gear pitch circle diameter Dro = 160 mm compressor speed at maximum engine speed Nc = 100,000 RPM turbine speed at maximum engine speed Nt = 100,000 RPM maximum engine speed Ne = 10,000 RPM speed ratio compressor to compressor pinion rc = 16 speed ratio turbine to turbine pinion rt = 4 speed ratio engine to planet carrier re = 1 Figure 2. shows the gear train arrangement for the above example in schematic form. The components shown are numbered as in Figure 1. with in addition, the engine to planet carrier drive system 7, the speed increasing gear drive 8 for the compressor, and the speed increasing gear drive 9 for the turbine.

When the shaft speed equation is calculated for the above values of the main parameters it reduces to the following equation : Nc = 40(we) - 3(Nt) For the compressor to reach its maximum speed of 100,000 RPM at an engine speed of 2,500 RPM, and thereafter run constantly at 100,000 RPM, at any engine speed between 2,500 and 10,000 RPM, the necessary turbine speed profile is defined by the above equation.

Figure 3. shows the calculated turbine speed profile in graphical form. The graph shows that the turbine must be held stationary until the engine reaches 2,500 RPM, and thereafter has to increase as a linear function of engine speed.

Figure 4. illustrates two further examples of compressor speed profiles, and shows the turbine speed profiles which would be necessary to generate them. In these examples the compressor profiles are of a smooth continuous form. The turbine speed profiles are therefore also of a smooth continuous form.

Within the practical limits of gear wheel sizes, and allowable gear wheel speeds, and the necessity of avoiding the compressor surge line, any desirable compressor speed profile may be generated, by suitable control of the turbine speed.

Any system required for controlling turbine speed, in order to achieve a desirable compressor speed profile over the useful operating speed range of the engine, is not part of this invention. It may be assumed for example, to be an additional function of an electronic, or computerised engine management system.

The management system would operate in conjunction with, a turbine brake system, and/or a turbine waste gate, and/or a turbine inlet restriction device, such as an inlet gas throttle, or a moveable inlet guide vane system. Alternatively, or additionally, a turbine discharge restriction device such as a discharge gas throttle or a moveable discharge guide vane system, could be used.

An additional function of such a management system could be to avoid compressor surge. That function of such a management system is also not part of this invention but may be assumed to be achieved by the use of a compressor charge air relief valve in the discharge line from the compressor, and/or a moveable discharge guide vane system in the compressor discharge diffusor, and/or a moveable inlet guide vane system at the compressor inlet.

Claims

1A differential drive system for centrifugal charge air compressors for internal combustion engines, in which the compressor speed, and therefore the boost pressure developed by the compressor, can be controlled throughout the useful engine speed range. The system uses an exhaust gas turbine to modulate the effective speed ratio between the engine and the compressor, to give control of the compressor speed.

2 A device as claimed in Claim 1, wherein the compressor speed and therefore the boost pressure is not controlled solely by the engine, as in a centrifugal charge air compressor driven solely by the engine.

3 A device as claimed in Claim 1, wherein the compressor speed and therefore the boost pressure is not controlled solely by the turbine, as in a turbocharger system having the compressor driven directly by the exhaust gas turbine.

4 A device as claimed in Claim 1 and Claim 3, wherein at low engine speeds, the compressor is able to be driven, and its speed controlled, solely or primarily by the engine, employing a drive system having a very high effective drive speed ratio, which allows the compressor to reach its maximum operating speed when the engine is rotating at relatively low speed.

5 A device as claimed in Claim 1 wherein the exhaust gas turbine is able to provide part of the power to drive the compressor.

6 A device as claimed in Claim 1, Claim 3 and Claim 5, wherein the compressor and exhaust gas turbine operating speeds can be selected independently, with consequent elimination of constraints on the optimisation of the size, weight, and efficiency of the compressor and of the exhaust gas turbine, when compared with a conventional turbocharger system.

7 A device as claimed in Claim 1 and Claim 3, wherein since the compressor boost pressure is maximised throughout the useful engine speed range, there is no delay between engine throttle demand and engine response, as may be the case with turbocharger systems in which the compressor is driven solely by the exhaust gas turbine.

8 A device as claimed in Claim 1 and Claim 5, wherein the exhaust gas turbine can augment the power produced by the engine, since the exhaust gas turbine is linked to the engine crankshaft by the differential drive system.

9 A device as claimed in Claim 1, Claim 5, and Claim 8 wherein if the compressor is partly or fully, externally braked, and/or if the incoming charge air is allowed to fully or partly bypass the compressor, the whole of the useful power developed by the exhaust gas turbine augments the power produced by the engine.

10 A device as claimed in Claim 1, wherein the intermediate speed shafts in the gear trains between engine and compressor, and which can be controlled to run at nominally constant speed throughout the useful engine speed range, may be employed to drive auxiliary engine components. For example, particularly, generators or alternators, but also hydraulic or pneumatic motors, and coolant pumps or cooling fans.

11 A device as claimed in Claim 1, and Claim 5, wherein the exhaust gas turbine could be replaced by a variable speed electric motor, or an hydraulic, pneumatic, or other prime mover if required, without affecting the compressor drive system and the benefits of that system as described in Claims 1, 2, 3, 4, and 7.

12 A device as claimed in Claim 1 and Claim 6, wherein the exhaust gas turbine can be smaller than in an equivalent conventional turbocharger, since the main means of boost pressure control is by turbine speed variation and not by diversion of the exhaust gas around a basically oversized turbine, as in a conventional turbocharger.

13 A device as claimed in Claim 1, Claim 6, Claim 8, and Claim 9, wherein the exhaust gas turbine size can be usefully made larger than in an equivalent conventional turbocharger, thus maximising the amount of exhaust gas energy which can be recovered in the turbine and used to augment the power produced by the engine.

14 A device substantially as described herein with reference to Figures 1 - 4 attached.