A new class of rotary piston machine suitable for compressors, pumps and internal combustion engines

COMBUSTION ENGINES GROUP

A NEW CLASS OF ROTARY PISTON MACHINE SUITABLE FOR COMPRESSORS, PUMPS AND

INTERNAL COMBUSTION ENGINES

J. M. Clarke, BSc, CEng, MIMechE* D. F. Walker, CEng, MIMechE* P. H. Hamilton*

The mechanisms and porting requirements of a new class of rotary piston machine are described. The machines can be regarded as a generalization to three dimensions of that class of planetary motion machine which includes the Wankel. Important features of the Wankel engine are retained, namely a compact arrangement, a sliding contact seal grid and multiple chambers on each rotor, but these new machines have twice as many chambers on each rotor, and they apply results of three-dimensional rigid body dynamics to select piston

motions which involve very low inertial forces. An experimental engine is described with some results.

1 INTRODUCTION PATENT LITERATURE contains hundreds of devices of great ingenuity which stem from a desire to avoid the reciprocating piston and yet connect chambers of varying volume with a rotating shaft. Only the Wankel engine has achieved any great commercial interest.

The machines outlined in this paper form a new group also based on compound rotation, i.e. rotation about a moving axis. But while the rotor axis of the Wankel engine sweeps a circular cylindrical surface the rotor axis in these machines sweeps a circular conical surface so that its centre of gravity, at the apex of the cone, does not move. These machines are not featured in surveys (1)-(3)t of rotary piston machines although Dr Felix Wankel was clearly aware (I) (4) that there were possibilities in this direction. The machines are free from vane or flap type components and they retain the very important ‘internal- axis’ feature which allows the adoption of a sliding continuous contact seal arrangement so that use as an engine is practicable as in the case of the Wankel engine and the Hamilton-Walker engine (5)$. In the latter the seal line is also a mechanical constraint on the motion of the piston. ‘External axis’ machines such as the helical screw compressor (a), the Dean gear engine (7) and the Marshall tri-dyne engine (8) suffer from their inherent inability to use contacting seals. Other important features of the Wankel such as the elegant use of ports to avoid valve mechanisms and the compact arrangement are shared by this new group.

It is demonstrated in Appendix 1 that the dynamics of the rotors in these machines are favourable. In some of them the inertial forces applied to the bearings are so small

This paper is intended for presentation at an Ordinary Meeting on 18th October 1972. The M S . was received on 8th May 1972 and accepted for publication on 21st July 1972. 22

* Turbo Machinery Department, National Gas Turbine Establish- ment, Pyestock, Farnborough, Hants. + References are given in Appendix 2.

$ No connection with the authors.

that it is possible to think in terms of the much higher speeds normally associated with turbomachines.

This paper outlines the mechanisms involved; the porting configurations suitable for various applications ; the sealing requirements and some of the experience obtained from tests of an experimental engine at the National Gas Turbine Establishment. Finally, some important differences between these machines and their parallel axis equivalents are outlined.

The motions and shapes are intrinsically three-dimensional and are unfamiliar to engineers. This makes their presentation difficult and explains the large number of illustrations in this paper. The machines themselves, however, are not inherently very difficult to mass produce.

1.1 Notation m,/m3, m2/m3 ratios of principal moments of

Effective area for leakage through seal gaps

Constant having dimensions of angular mo-

Couple magnitude. Functions of and /3. Function of Ap/p. Shaft moments of inertia about axes of rota-

Mass of rotor. Principal rotor moments of inertia. Initial fluid mass in chamber. Mass of fluid leaked through seals. Defined by equation (15), for a machine with

fixed ports it is an integer greater than 1 and equals the number of chambers on each side of a rotor.

Fraction of cycle period for which outflow can occur from chamber.

Pressure in a chamber.

inertia.

and clearances.

mentum.

tion.

Proc lnstn Mech Engrs 1972 Vol 186 62/72

744 J. M. CLARKE, D. F. WALKER AND P. H. HAMILTON

AP

R T

V v c

B

t

R

P Pc 7

4 *

Pressure drop across seals, positive for outflow from chamber

Gas constant or tip radius. Absolute temperature. Time. Mean speed of rotor tip. Initial volume of chamber. Angle between shafts in the Hooke’s coupling. Eulerian angle between rotor polar axis and

fixed direction of maximum angular momentum, i.e. in precessing mechanism the angle between the rotor and mainshaft axes.

Density of fluid in chamber. Initial density of fluid in chamber. Cycle period. Eulerian angle defining rotation of plane con-

taining 0 about fixed (mainshaft) direction. Eulerian angle defining rotation of rotor

about its polar axis relative to plane containing %, it is zero when principal axis m2 is normal to plane containing 8.

2 MECHANISMS In the discussion in Appendix 1 the three mechanisms outlined in Figs 1, 2 and 3 can be described as arrangements for connecting compound rotations of a ‘near free body’ type, i.e. motions involving low inertial forces, with simple rotation of a shaft. The reasons for this are discussed in Appendix 1.

2.1 Precession In Fig. 1 the motion of the rotor is a precession. Rotation of the mainshaft at a rate 6 causes simultaneous rotation of the rotor at a rate 4 relative to the tilted portion of the mainshaft. The ratio between the speeds of the rotor and

Rotor

Bevel shaft \

-. U

,- % Layshaft-

b

Bevel shaft\ n-/ Rotor

Mainshaft

-

Layshaft

a Two shaft revolutions for each rotor revolution. b Three shaft revolutions for each rotor revolution.

Fig. 1. Precessing rotor mechanisms

shaft is controlled by the bevel and layshaft gears. The two examples show gear proportions appropriate to the fre- quency ratios of most practical interest. The case where = -24 leads to applications involving one cycle of

volume variation for each rotor revolution; the case where 6 = -3$/2 leads to applications such as the four cycle engine in which two cycles of volume variation occur during each rotor revolution. Many alternative arrangements of gears involving different bevel shaft and layshaft rotation rates are possible. More compact arrangements can be made using sun and planet gears instead of a layshaft. The straight-through mainshaft and large bearing areas give a robust mechanism suitable for application in high pressure machines such as internal combustion engines.

2.2 Hooke’s coupling Fig. 2 is the familiar Hooke’s coupling. In this case the central member (rotor) is outside the coupling and the shafts are constrained by fixed bearings with intersecting but angled centre lines. It has the disadvantage compared with the precessing mechanism that the shafts and trunnion bearings are overhung and the bearings are

-Rotor

Fig. 2. Hooke’s coupling mechanism

Slidi

! I

I i

I I I ! ‘. \

‘-- Contact face .. k u F Rotor

Fig. 3. Sliding apex mechanism


ROTARY PISTON MACHINE SUITABLE FOR COMPRESSORS, PUMPS AND I.C. ENGINES 745

Fig. 4. Precessing rotor with four working chambers

necessarily limited for space. However, it avoids the use of gears. The motion of the rotor resembles that of precession with $ = -24 and the possible applications overlap with that case.

ment devices by providing means for altering the angles. In the former the angle between the shafts is variable and in the latter the angle between the shaft and the normal to the face F is variable.

2.3 Sliding apex The third mechanism, Fig. 3, causes the rotor to have exactly the same motion as in the Hooke’s coupling but the constraint of the second shaft is replaced by sliding contact between the flat circular face, F, and the cylindrical apex, C, of the rotor. This contact is practical in some cases because the inertial loads transmitted through the contact can be made small and fluid pressures in the two chambers exert no extra loads at that point. The avoidance of a second shaft and hub sphere clearly simplifies the construction enormously when compared with that of Fig. 2. Both the Hooke’s coupling and the sliding apex mechanisms can be incorporated into variable displace-

3 SEAL SURFACE SHAPES It is shown in Appendix 1 that each of the above mechanisms constrains the rotor to move in a manner which satisfies the condition for making a positive displacement device with ports in a static casing. The shape of the casing is defined by fixing a seal line in the rotor and then moving the mechanism. As the rotor moves this line sweeps a surface which is continuous. In the case of the Wankel engine the casing surface is an epitrochoidal cylinder. The remainder of the rotor outline can then be defined as the largest rigid object which does not touch this surface. Figs 4 and 5 show computer generated pictures of precessing rotors moving relative to fixed static surfaces

Fig. 5. Precessing rotor with six working chambers



defined in this way. In the particular cases drawn the apex seal lines have been chosen straight and the apex seals on each side lie in a plane. The two sides are the same shapes. Four separate chambers are confined between the rotor and the static surfaces in Fig. 4 and six in Fig. 5. The spherical surface round the outside has been omitted to show the rotor at various positions in its cycle. A rather surprising result evident from Fig. 4 is that the stator seal surfaces appear to be flat. They are not exactly flat but the difference causes seal movements small enough to be accommodated by a normal seal grid so that in practice it is not necessary to generate seal sliding surfaces any more complex than flat and spherical. The faces of the rotor can also be made flat.

The above discussion has been based on the use of ‘line’ seals in the rotor. An alternative is to define ‘surface’ seals in the rotor representing for instance a cylindrical surface on the rubbing nose of an apex seal. In the case of the Hooke’s coupling and sliding apex mechanisms cylindrical apex seal surfaces on the rotor having axes concentric with a pair of trunnion bearings define exactly flat surfaces for the casing.

0 120 240 360SHAFT480 600 120

Fig. 6. Two-cycle engine

4 PORTING Ports for fluid to enter and leave the various chambers at appropriate times can be positioned in the outer spherical surface. Figs 6-9 show projections (geographically they would be known as Mercator’s projections) of this spherical surface on a cylinder. The position of the rotor against this surface is shown at intervals during a rotor revolution. It can be seen from these schematic diagrams that movement of the rotor periphery over the surface can be used to open and close suitably positioned ports for the four most important working cycles. This ability of the rotary piston to avoid valve mechanisms, camshafts, etc., is one of its most attractive and economic features. The only one of these cycles for which a reciprocating piston can achieve the same effect is the two-cycle engine and even in this case more advantageous timing can be achieved with the rotary piston. For instance the exhaust port may be opened and closed before the inlet port.

Because several chambers use the same ports there is greater continuity of flow through them than can be obtained with reciprocating machines which need multi- branched manifolds. For example, in each case, the single inlet port is never closed. The only cycle for which the porting requirements have compromised the rotor shape

1

2

60 120 180 240 300 360 ROTOR

0

Fig. 8. Sliding apex pump

3

1 4

Fig . 7. Four-cycle engine

Proc lnstn Mech Engrs 1972

- __ ~ _ c - _c . - - -- .. Volume ,pK-\. Inlet area “-/ Outlet area

1 4‘ \<,’ /!-\ \/f\h*.

120 I80 240 300. 360 ROTOR

Fig. 9. Compressor

->A 60 0

Vol 186 62/72


is the compressor-expander where a recess in the rotor is used to communicate with the high pressure port and the width of the rotor at its tip has been increased to allow room for that port.

5 THRUST L O A D S A N D BALANCING All the machines present surfaces to the fluid which lead to thrust loads on the shafts. In systems using more than one rotor these loads may be balanced within the shaft system, otherwise substantial thrust bearing surfaces are needed. Symmetry of a rotor ensures that there need be no time- average thrust even with a single rotor. The alternating nature of the load allows the powerful squeeze film effect to be used on the load bearing surfaces.

The mechanisms were all chosen to comply with a requirement for low inertia loads between the rotor and its mechanical constraints. Nevertheless, the constraints do experience some loads and it is desirable that these are balanced within the machine and not transmitted to its mountings.

There is a considerable difference between the balancing of the processing and Hooke’s coupling mechanisms. For the former the small out-of-balance couple can be com- pensated entirely by weights on the mainshaft but for the latter there is no shaft rotating at twice rotor speed and this simple solution is not possible. One answer in this case is a second rotor or its dynamical equivalent; a second rotor also balances thrust loads within the rotating system.

6 S E A L I N G It is not practical to give geometric details of the various seal configurations appropriate to the many different applications. The analysis of this section provides a relatively simple figure of merit for any positive displacement machine and shows how this may be compared with the requirements of the cycle and fluid for which the machine is intended. It allows, for instance, a rational basis for deciding whether it is practicable to use close- clearance ‘sealing’ as opposed to the more expensive contacting grid seals. So many machines have failed in the past because their design did not recognize the magnitude of the sealing problem that it seems probable that this analysis will also be of interest outside the subject of this particular paper.

As a criterion it is assumed that the mass flow leaked should not exceed 5 per cent of the mass flow of the per- fectly sealed machine. Furthermore, for analytical simplicity it is assumed that the leakage does not significantly alter the cycle pressures and temperatures. A more accurate and complex method of analysis has been developed at N.G.T.E. for assessing the effects of leakage but a description of it is outside the scope of this paper and its results are not significantly better from a designer’s point of view.

The mass flow rate mL for a fluid leaking through an effective leakage area A, from a chamber at pressure p and density p is given by

n i L = A L d G f e ) . . . where A p is the pressure drop across the seal and f is a function of Aplp.

Integrating for the fraction n of the cycle period T for

Proc lnstn Mech Engrs 1972

which A p is significantly positive and assuming A , is constant during the cycle gives

as the mass lost from a chamber during a cycle.

mately

where rn, is the trapped mass; V, is the initial chamber volume and p , is an appropriate initial charge density.

The 5 per cent leakage criterion may be written

The fresh fluid held initially by the chamber is approxi-

m, = p , V c . . . . (3)

rn, > 20m, . . . . (4) and using equations (2) and (3) this can be expressed in the form

The terms collected on the left give a figure of merit for a machine; they depend on the size, V,, speed, T , and seal standard, A,, for a particular machine. The terms on the right-hand side (r.h.s.) can be evaluated for a particular thermodynamic cycle of an engine or pump. Inequality (5) therefore provides a criterion by which a particular machine design may be compared with the requirehents of its cycle.

In the case of two particularly important cycles the right-hand side may be simplified and approximated further. Firstly, in the case of internal combustion engines or high pressure compressors the leakage is mainly through choked flows so that for a perfect gas with a specific heat ratio of 1.4

P R T p = - and f = 0.6847 . . (6)

so that inequality (5) becomes V, 20~0.6847J’ p (:) &’ dRp, o T T d - * (7)

For a four-diesel engine the r.h.s. of inequality (7) is typically 17 000 m/s and for a petrol engine more like 5600 m/s.

Secondly, for low pressure blowers or liquid pumps where compressibility is negligible

p = p , and f = i y . . (8)

furthermore A p is almost constant and n N 4 so that inequality (5) becomes

For an atmospheric blower with d p = 10 kN/m2 and pc = 1.21 kg/m2 this gives a r.h.s. of 1285 m/s and for a water pump with A p = 1 MN/m2 and p c = 1000 kg/m3 the r.h.s. is 448 m/s.

The above analysis illustrates for instance that in any positive displacement machine with a 500 cm3 chamber and a 3000 rev/min shaft speed leakage areas less than 0.75 mm2, 2.2 mm2, 19 mm2 and 56 mm2 are needed for the diesel, petrol, blower and liquid pump respectively to avoid prohibitive performance deficiencies arising from leakage.

Vol 186 62/72


Two-cycle engine 1 N.G.T.E.

Four-cycle engine

Table 1

Not Not suitable suitable

~ ~~ ~

cycle I Mechanism

Liquid p v p - motor or zur blower

Compressor- expander

Precessing 1 Hooke’s 1 Sili i coupling

N.E.L.

7 A TWO-CYCLE EXPERIMENTAL UNIT By combining different porting arrangements with different mechanisms many different machines can be con- structed. The possibilities are conveniently identified in Table 1.

The National Engineering Laboratory is examining a

Rotor

sliding apex blower. Effort at N.G.T.E. has concentrated on the precessing machine in the form of an experimental two-cycle rig supplied with scavenge air from the per- manent test facilities. Fig. 10 shows a longitudinal section and Fig. 11 shows a part-sectioned model. The 254 mm (10 in) diameter rotor forms four working chambers each of 510 cm3 (31.1 in3) swept capacity. The rotor and casings are cast in S G iron with integral cooling passages for oil in the rotor and water in the casings. Rotor bearing lubricating oil and rotor cooling oil are supplied by separate drillings in the mainshaft. If it were successfully developed as a turbocharged compression ignition unit with a b.m.e.p. (brake mean effective pressure) of (say) 1400 kN/m2 (203 lb,’in2) such a unit would deliver 118 kW (160 b.h.p.) at a rotor speed of 2500 rev/min. For application as a compression ignition engine a separate chamber was provided for combustion on the lines of the Ricardo Comet.

The design rotor speed was limited to 2500 revlmin in order to stay within normal limits on injector pump speeds and seal sliding speeds. These restrictions mean

cooling oil

Lay shaft

Ports’

Fig. 10. Longitudinal section through the experimental two-cycle engine

‘Mainshaft

Fig. 11. Part-sectioned view of the experimental engine

Proc lnstn Mech Engrs 1972 Vol ‘I 86 62/72


%crion shown in fJgures 13 and 14 .Apex real

a Rotor close to casing. b Rotor away from casing.

Fig. 14. Section through third design of apex seal

pox seal -/

Fig. 12. Seal configuration

that this particular rig did not exploit the potential for high speed operation and, at the same time, it made heavy demands on the standard of sealing required.

7.1 Seals Fig. 12 shows the seal grid in schematic form to illustrate the names used for the various seals and the position of the enlarged section shown in Figs 13 and 14. The sealing of each chamber is made up from four pieces.

(1) A leading apex seal. (2) Half the circumference of the hub ring. (3) A trailing apex seal. (4) A tip seal.

The two apex seals are the least conventional because, like the apex seals of the Wankel engine, they tilt as they slide and are therefore limited to line contact as opposed to

r Apex seal

osing

Apex seal

Rotor

asing

Principal leakage areas

a Rotor close to casing. b Rotor away from casing.

Fig. 13. Section through first design of apex seal

Prac I nstn Mech Engrs 1972

surface contact. Also, like the Wankel the seal grid contains four corners. The tip seal is semicircular like half a piston ring in the rotor. The circular hub seal is located in the casing, and seals by inward acting forces as it slides against the rotor hub sphere. The apex seals are located by grooves in the rotor and slide on the casings. All the seals are spring-loaded to ensure correct seating for start- ing but are mainly loaded by gas pressure in the conventional manner. The tip seal has not changed much during development and has been made from a piston ring grade of cast iron. Hub seals were originally cast iron but more recently have been made in En 31 steel. Apex seals have been tried in several materials including carbon and cast iron.

Fig. 13 is a section through the apex seal of the first design. Radial support of the apex seal is on the rotor. The components are drawn in the extreme positions they may occupy during operation. This movement between extremes can arise from a combination of bearing clearances, different thermal expansions, load deflections, surface form and position errors and timing gear backlash. In the first design, Fig. 13a, when the gas pressure forces the rotor away from its casing leakage areas of up to 6 mm2 can be opened up at each apex seal.

The third design is shown in a similar fashion in Fig. 14. The apex seal is now supported radially by sliding contact on the casing and is made smaller and lighter. This simplifies the corner problems between the tip and apex seals. The hub ring secondary sealing surface has been moved out of the locating groove so that the latter no longer provides a circumferential leakage path. The part of the hub ring on the expansion side of the engine has been provided with an integral water-cooling passage and its expansion gap has been sealed with a stepped design. The three- piece design of apex seal can accommodate changes of seal-line length by relative motion of its members. The maximum area exposed by clearance allowances is now 0.5 mm2 and at the most important period when gas pressure forces the rotor away from the casings the corner leakage disappears. Many other detail differences not apparent in the diagrams have been introduced. In particular the surface finish and form have been much improved by lapping. All surfaces are either flat or spherical so that they lend themselves to this kind of precision finishing operation.

Vol 186 62/72


Identified

P I First design before modifications -

7.2 Test performance At the time of its design the crucial nature of the sealing was not adequately recognized. The analysis in Section 6 is hindsight. The first tests, however, were disappointing and the engine would not drive itself in spite of using petrol and spark ignition. The only consolation was that the mechanism ran remarkably smoothly. The trouble was diagnosed as a mixture of poor combustion and poor sealing in proportions depending on the mood of the moment. Subsequent analysis, however, showed sealing to be the predominant problem. Pressure diagrams during motoring tests were analysed to establish leakage levels which were expressed in terms of effective leakage area, mm2. A simpler static method of testing leakage was also used in which with one chamber at t.d.c. (top dead centre) the pressure in the chamber was maintained by supplying measured quantities of compressed air through a sparking plug hole. The quantity of air needed gives a direct indication of the leakage. This static method of testing the seals has proved invaluable and correspondence between the leakage areas estimated in this way and by analysis of motoring results has been good. Fig. 15 shows the progress made with the sealing effectiveness first by detail modifications of the original design and then by two new designs. The first of these new designs did not live up to expectations, particularly at higher speeds, but the second appears to have provided the sort of improvement necessary. The method of measuring seal performance by analysis of motored pressure diagrams becomes inaccurate for leakage areas of the order of 1 mm2 at the higher speeds.

Remembering that there are two seal lines on each element there are 1500 mm of seal line round each chamber (twice the length including valve seatings for a reciprocating piston of the same capacity). The initial seal standard (21 mm2) therefore represents an average gap of 0.014 mm (0.0006 in) and the current average gap at speed is about 0.0015 mm (0.000 06 in). The analysis does not apply accurately to such small clearance but nevertheless it is obvious from these figures that any machine adopting a non-contacting ‘seal’ design, say 0.01 mm clearance, would be inadequate for use as an engine.

Measured chamber pressure diagrams and some results of a computer simulated analysis of the pressures are compared in Fig. 16. The upper graph shows clearly the shortcomings of the original seal performance and the middle graph shows the current standard. At the time of the first tests, rig limitations prevented the use of rotor

LEAKAGE AREA mmz 0 -

First design Identified before modifications

after First design rnodificotions 1 -/ed

First design after rnodificotions

Second design

Third design

Fig. 15. Seal effectiveness improvement

631 compratio

4 8 0 120 -60 0 60 120 1 ROTOR ANGLE dqnc

0

a Motoring test on first seal design. b Motoring test on third seal design. c Power test on third seal design.

Fig. 16. Measured and simulated pressure diagrams

motoring speeds above 750 rev/min so that there is no direct comparison possible with the higher speed motoring results. The lower graph represents the best torque achieved so far by using petrol and spark ignition. The indicated mean effective pressure is 400 kN/m2 which is still low due to leakage and the measured brake power at these conditions was 18-1 kW (24.3 b.h.p.). This corresponds to an average b.m.e.p. of 300 kN/m2 for the four chambers and a mechanical drag of about 100 kN/m2.

8 COMPARISON WITH OTHER ENGINES The principal advantages of rotary piston machines over their reciprocating equivalents are their smaller size and greater smoothness. The rotary piston machines should in the long term reduce costs because material costs are a high proportion of the cost of mass produced mechanical items. The latter is the direct result of the improved balancing and the removal of valve gear. Clearly, if seal materials develop sufficiently to ensure satisfactory life, the long term prospect for the rotary engine is good. In the short term, however, if manufacturers are forced to write off a part of their enormous investment in reciprocating machines it will be the result of other pressures such as the need in cars for more underbonnet space to accommodate air-conditioning, anti-pollution and extra silencing, or in lorries the saving of cab space in the face of increasing installed horsepower requirements.

The most highly developed rotary piston machine to


ROTARY PISTON MACHINE SUITABLE FOR COMPRESSORS, PUMPS AND I.C. ENGINES 75 1

date is the Wankel planetary motion engine. Significant technical differences between this and the compound rotation machines discussed here include:

(1) The inertial loads in the Wankel engine with a stationary trochoid* are of similar magnitude to those in a reciprocating piston engine of the same power, shaft speed and firing impulses per revolution. The question of inertial loads in this new family of compound rotation units is discussed in Appendix 1 but they are very much smaller so that these new engines should be capable of development to higher speeds.

(2) The new compound rotation units can use both faces of the rotor and therefore have twice as many working chambers per rotor. (3) The geometry of the trochoid and the need to pro-

vide a transfer passage in the rotor flank places an upper limit on the feasible compression ratio in the Wankel engine. This limit is very much higher in the precessing rotor versions so that it should be possible to achieve compression ignition in one stage.

(4) The timing gears are more complex in the precessing machines but by way of compensation they do not impose a limit on the mainshaft diameter as they do in the Wankel engine.

(5 ) The porting arrangements outlined in Figs 6 to 9 employ single inlet and outlet ports communicating with from two to six chambers. This ensures much greater continuity of flow than can be achieved in a conventional manifold arrangement. This is an advantage from the point of view of mixture distribution in carburetted engines and it helps maintain reactor temperatures in exhaust emission control devices. Peripheral ports, i.e. ports in the trochoid, on the Wankel engine are said to be suited to high speeds but have too much overlap for opti- mum low-speed running. Side ports are less suited to high speeds because they involve sharp turns in the flow. The corresponding precessing rotor machine can use a port arrangement like that shown in Fig. 7. It features large port areas with no flow turning and no overlap. * The original Wankel engine used a moving trochoid and avoided

inertial loads but the complications inherent in moving ports led to the adoption of the present configuration.

U

a Precessing rotor. b Wankel.

Fig. 17. Comparison of combustion chamber shapes

(6) In the four cycle engines there is another considerable difference in the combustion chamber airflow near t.d.c. Fig. 17 shows successive combustion chamber shapes for both engines from which it can be seen that the rapid transfer of gas across the waisted portion of the Wankel engine chamber is absent from the precessing version and so is the persistence near the trailing apex seal of a pocket of gas which is remote from the combustion process until late in the cycle. These differences may help reduce heat loss and exhaust emissions when compared with the Wankel engine. (7) In the two-cycle casing machines the surfaces can

be made without special generating or cam-following machinery and where flat and spherical surfaces are used a high degree of precision and good finish is obtainable by lapping.

9 CONCLUSION Attention is drawn to a new class of rotary piston machines for the exchange of fluid energy with a rotating shaft. The machines are characterized by the near free body motion (precession in some instances) of their working elements and a compact arrangement with several chambers on each rotor. They offer a unique combination of freedom from inertial loads, the ability to seal by continuous sliding contact and freedom from valve mechanisms with a stationary casing containing-inlet and outlet ports.

One of these machines has been built and tested as a two-cycle engine and extensive effort has been necessary to improve the sealing effectiveness. However, recent results have pointed the way to a solution of the seal effectiveness problem for certain compressor duties and a high speed engine. A much more extensive programme of work is necessary to improve seal life and to satisfy the effectiveness requirements of low speed engines but it seems probable that improvements can be made if the development effort is forthcoming.

There are similarities with the Wankel engine but there are also significant differences many of which favour the new machines.

10 ACKNOWLEDGEMENTS The authors wish to acknowledge with gratitude the efforts of many of their colleagues at N.G.T.E. who con- tributed to this work and also the staff of the Computer Aided Design Centre at Cambridge for the use of their facilities to produce Figs 4 and 5. This paper is published by permission of the director of the National Gas Turbine Establishment and is Crown Copyright Reserved.

Fig. 18. Gimbals mechanism and Euler’s angles



APPENDIX 1 I N E R T I A L FORCES FOR C O N S T R A I N E D PRECESSION,

H O O K E ' S C O U P L I N G M O T I O N A N D T H E W A N K E L E N G I N E

Precession and free-body motion Fig; 18 shows gimbals arranged to allow three independent rotations of a rigid body or rotor about its centre of gravity. Rotations measured on each axis correspond to changes in the Eulerian angles 0, J, and 4. If no frictional or other forces act on the rotor and the inertia of the frames is negligible the body will move freely (9) so that

sin $ cos J, sin 6' . . (10) m1-m2 B = a - 11111712

1 cos2 $ sin2 J, m3 m , m2

t j = a (-----) cos 8 . (11)

= (cos2 J, sin2 $1 a - +- . . ,

m, m2 The principal moments of inertia are m,, m2 and m3 and m3 corresponds to the polar ($) axis and m2 to the 6' axis when J, is zero. The multiplier, a, is a constant of the motion having dimensions of angular momentum. Simple rotation is the case when 6' = 0 because then

e = 0 and $+$ = p / m , . . (13) Steady precession occurs when m, = m2 because then

e = O and $ =cos6' --1 4 . (14) c: For rotors with repeating geometry such as those shown

in the text m, = m2 so that for this class of rigid bodies free-body motion is steady precession.

Condition f o r a positive displacement machine in a $xed housing The Wankel engine achieves its sealing using a three-fold repeat of rotor geometry with an eccentric shaft which rotates three times for each rotor revolution. In the more general case a rotor with an N-fold repeat of geometry moving cyclically so that 0 returns to its initial value and J, decreases by 2 4 N - l ) /N in the interval needed for 4 to increase by 2~ can be sealed in an analogous manner.

In the case of precession the constraints are such that N- 1

N 8 = 0 and * = --$ . . (15)

Equations (14) and (15) cannot be satisfied simultaneously because N is a positive integer greater than 1, cos 6' < 1 and m,/m3 > +. The closest approach to free-body motion is achieved for N = 2, 0 small and mJm3 N so that the rotor should be thin relative to its diameter.

In the case of Hooke's coupling the Eulerian angles satisfy

. . . (16) - 112

1-sin2gsin2 4) ) (17)

where /3 is the angle between the shafts, and the axis of #J is inclined at /3/2 to both shafts.

An increase of 277 in 4 returns 6' to its initial value and reduces 4 by T so that this mechanism permits sealing of a rotor having a two-fold repeat of geometry.

The precession with N = 2 and the motion of the Hooke's coupling are very similar. For /3 = 20' and the same values of #J the maximum difference is 0.22 in # and 0.16 in 6'.

Both precession with N = 2 and the motion of the Hooke's coupling therefore satisfy the requirements for sealing of the Wankel engine type but are also near free- body so that inertial loads are low.

Inertial couple for steady precession The only external force applied through the bearings to balance the rotor inertia takes the form of a constant couple rotating in phase with the shaft (and 4) having a magnitude C given by (9) and (10)

-- C - N s i n O(l-N(l-cos O(1-2))) (18) m,Q2

where Q is the apparent rotor rotation rate defined by Q = d+$ and 6' is the fixed tilt angle of the shaft. This formula holds for non-integer values of N although only integer values from 2 upwards permit positive displacement machines in stationary casings.

Fig. 19 shows contours of constant couple covering the range of positive N values from simple rotation (N = 0) through the swash plate ( N = 1) to the wobble plate ( N = a).

Inertial couple for Hooke's coupling motion The shafts are inclined at angle ,6 to each other and the axis of 4 coincides with one shaft. We assume first that the motion of this shaft is known so that it can be written

where 0 is a constant and F(4) is a known function of 4. The Langrangian equations of motion can then be used to give the couples C, and C,,, conjugate to the Eulerian angles 6' and 4. They are

$ = .Q(F{#J})-'" . . . (19)

. . . (20)

PRECESSION RATE^+^-, Fig. 19. Contours for constant couple for precessing

motion

Proc lnstn Mech Engrs 1972 Vol 7 86 62/72


and c* G 2 tan /3 cos +

m3Q2 F 1+tan2 ,6 cos2 w=-

(21) d F l + B tan2 /3 cos2 + d+ 2F2( 1 + tan2 /3 cos2 +)

--

where G = tan /3 sin q5/( 1 + tan2 ,6 cos2 +), A = ml/m3 and B = m2/m3.

It will normally be more convenient for bearing calculations to express these results relative to the shafts. If C, is the couple having an axis parallel to that of + and the first shaft and C B is the couple having an axis normal both to the second shaft axis and its trunnion bearing axis (couple C, controls the normal force on the apex of the sliding apex machine), then

CA = C,+Co sin + tan ,!?/(I +tan2 /3 cos2 +) (22)

. (23) C, = - C,/(cos ,6( 1 + tan2 /3 cos2 +)1’2)

One simple case arises when the first shaft is forced to rotate at constant velocity then F = 1. A second case corresponds to free rotation of both shafts then CA = 0, or equivalently the kinetic energy is constant and

A tan2 ,!3 sin2 + 1 + tan2 /3 cos2 + +Btan2,!3cos2++1

(1 + tan2 ,6 cos2 +) F =

Some values of C, and C, are shown in Fig. 20. The preceding results can be used with shafts of finite

inertia by making adjustment to the rotor inertia. If the shaft inertias are Z A and I , for the first and second shafts then the ‘equivalent rotor’ has principal moments of inertia given by

m: = m,+Z, . . . . (25) m i = m 2 + Z A . . . * (26) m i = m 3 + z A + I B . * * (27)

Comparison with the Wankel engine inertia forces A typical Wankel engine with rotor mass M, rotor radius R and shaft eccentricity R/7 experiences an inertial force between rotor and shaft of magnitude

where V is the mean apex seal velocity. This formula can be derived as a special case in the limit of equation (18) or by considering the circular motion of the rotor centre of gravity.

A four-cycle ( N = 3) precessing rotor engine of similar power would have a rotor of similar mass My tip speed V and radius R. It would have a value for ml/ma of about 13/24 (corresponding to the proportions of a disc with thickness equal to half its radius) and a value for 6’ of about 10”. The couple magnitude from equation (18) can be expressed as

ICl = 0.1682MV2 . . * (29)

I 1 I I I I I

t -0.06

I I I I I I

0 30 60 90 120 150 180 SHAFT ANGLE $

-Constant shaft speed ------ Free rotation

p = 200

Fig. 20. Couple for Hooke’s coupling motion

If this is supported by bearing surfaces a distance R apart the load on each is

MV2 P, = 0.1682- . (30) R * .

Comparing equations (28) and (30) shows that for similar rotor masses and tip speeds the Wankel engine inertial forces exceed those of the precessing rotor engine by a factor of about 7. The corresponding factor for N = 2 is about 17.

APPENDIX 2 R E F E R E N C E S

(I) WANKEL, F. Rotary piston machines 1965 (Iliffe Books Ltd,

(2) SISTO, F. ‘Comparison of some rotary piston engines’,

(3) CHINITZ, W. ‘Rotary engines’, Scienr. Am. , February

(4) WANKEL, F. ‘A slant-shaft rotary piston engine’, U.K.

(5) ‘The Walker rotary’, Motor, February 1969. (6) WICHERT, K. W.

London).

770B, SAE National Powerplant Meeting, October 1963.

1969.

Patent Specification No. 805370, December 1954.

‘Characteristics of helical, rotary, positive displacement compressors’, A.S.M.E. 61-MYD-18, May 1961.

(7) DEAN, W. C. ‘A new rotary piston engine’, Mech. Engng, October 1964.

(8) ‘Marshall tri-dyne rotary engine’, The Engineer, March 1968. (9) WHITTAKER, E. T. AnaZytica2 dynamics 1944 (Dover Publi-

(10) THOMSON, W. T. Introduction to space dynamics 1961 cations Inc., New York).

(John Wiley and Sons, New York and London).


D261

Discussion

A. J. S. Baker Member I should like to make a few comments regarding the mechanical specifics of this very unusual class of machine. Like the Wankel epitrochoid machine, the torque is not transferred between the rotor and crank via the gearing. However, the gearing plays an important part in keeping the rotor accurately phased; in doing so it is subject to significant loadings due to inertial forces induced by torque fluctuations. These forces are usually of short duration and reverse during the cycle. It is difficult to compute the precise value of the gear loadings instantaneously as they are influenced by several dynamic components. T o contain this, the gearing may need to be more generously designed than would at first be supposed if scuffing and other difficulties are to be avoided. The drive requirements are somewhat similar to the zero power requirements of dynamic balancers used in reciprocating engines, which, in fact, sometimes need a drive capacity close to the average engine torque.

The Clarke engine offers better scope for combustion chamber design than some other rotary machines, since it does not suffer from parasitic volumes at minimum combustion volume. However, in the current examples on show, it would appear that cooling around the chamber may be inhibited by the thinness of the rotor section. Could the authors state whether there is any geometric objection to making the basic rotor rather thicker at the section occupied by the chamber ?

Examination of the crank-displacement diagram shows this to be sinusoidal. However, first approximation calculations suggest that this is not quite so. Could the authors define the divergence ?

The machines promise exceptionally high volumetric efficiency since the flow through them is pulsatory but virtually continuous. In these circumstances, unit heat fluxes for given gas conditions should be relatively low owing to the absence of turbulence. Many data are now available on the heat fluxes at the sort of fluid pressures and temperatures to be expected for the Clarke-Hamilton- Walker engine. Had the authors considered any comparisons with the relatively large swept areas of their machines ? Some data from current Wankel epitrochoid machines suggest that they at least may be somewhat critical in this direction.

It would appear that the length of seals subject to lubrication by constant loss methods is somewhat higher than that of the Wankel and significantly more than an engine with reciprocating pistons. In the prototype engine it appeared that the opportunity to make the circumferential end seals recirculate oil had been avoided. Could the authors give their comments on this point which appears of importance to pollution and chamber fouling ?

The use of lubricant on a constant loss or sacrificial basis appears at first sight a retrograde step and the oil industry might be more interested in developing future oils which could be recirculated rather than a new breed to be used sacrificially.

Continuing on the circumferential end seals, it would be interesting to see any in-cycle pressure diagrams taken between the rings. The seals represented an application of the normal Ramsbottom piston ring pack which depends upon static and dynamic leakage reaching the between- ring spaces sufficiently quickly in the cycle to pressurize these spaces and hence partly relieve the sealing force acting upon the first ring (11). In the case of the Clarke machine, only part of the first ring is subject to pressure at any time; thus before any between-ring pressure could accumulate, gas from the pressurized segment would have to pass around the between-ring annulus and in a normal cylinder pressure diagram this delay would result in considerable pressure reversals within the pack. Could it be that dynamic leakage between segments had been partially responsible for some loss of performance ?

A further problem of using rings in this manner is the effect of wrapping such as that used in a capstan. Clearly the ring would be prevented from rotating with the rotor under pressure, as the total side forces and areas exceed the radials. In these circumstances it depended upon the circumferential position of the ring ends in respect of maximum load on the ring as to whether the ring might wrap or unwrap on the rotor. I n either case, the effect might be to reduce sealing.

In connection with these rings, it was interesting to note that they were made of En 31, and this reminded one of the late R. C. Cross who pioneered rolled steel rings of this material with great success. In particular, Cross rings never break and this might explain why they had been chosen in this application.

Finally, rather than attempt to find a name for this type of engine which might satisfy an academic minority and mean very little to practical engineers, it would be better to call this new class simply Clarke machines. Unlike Wankel, Clarke had not apparently considered other types of rotary, thus the description would be quite specific.

REFERENCE

(11) BAKER, A. J. S., CASALE, P. G. and SLOAN, H. ‘Piston ring loading factors and a method of wear measurement in engine cylinders’, 9th CIMAC Congress 1971, paper No. A.29.

S. A. Egberongbe Guildford Rotary piston machines are not a novelty, but because of

Vol 186 62/72 Proc lnstn Mech Engrs 1972

D262 DISCUSSION ON J. M. CLARKE, D. F. WALKER AND I?. H. HAMILTON

4 Chamber N.G.T.E. 2 Rotor 2 htre Wankel P = = f + E 4 Cylinder 2 litre reciprocating

an almost infinite variety of possible arrangements, both of moving components and power taken off, they have confused engineers as to where to begin. The adoption of precision motion, resulting in low inertia forces and con- sequent high speed capability, gives this new type of rotary machine an advantage over all others.

The use of integral cooling passages for oil in the rotor and water in the casing appears inadequate for an engine with high power/weight ratio. Could the authors give an indication of the magnitude of the thermal loading at a rotor speed of 2500 rev/min when the unit delivers 118 kW? Could the authors also explain how they hope to avoid the risk of oil seeping into the combustion chamber.

2850 32 2200 85 m* 54

F. Feller Fellow I found the introduction, which establishes the category into which the new machines can be placed, most helpful. Both the engine commonly referred to as the Wankel engine and the group of engines described in the paper have precessing rotors and internal axes. Why not then refer to the former as the ‘parallel axis’ rotary engine and to the latter as the ‘slant axis’ rotary engine ?

Of particular interest to me was the ‘Comparison with other engines., I agree completely with the summary of rotary engine virtues, vis-a-vis the reciprocating engine, but I am not convinced that any of the differences between the ‘slant axis’ rotary engine and the ‘parallel axis’ rotary engine give the former an overriding advantage in practice, for instance:

(1) The ‘parallel axis’ rotary engine has proved itself to be a very smooth engine capable of running up to high speeds. In fact the only limitation to speed has been apex seal wear and the rate of combustion. How could inertia loads help in this respect ?

(2) Having twice as many working chambers per rotor could become a thermal problem and hence a disadvantage.

(3) T o be able to have a higher compression ratio and a more compact combustion chamber is attractive for a diesel version, but I would set against this advantage the need for a very large end-thrust assembly and the more difficult gas seal configuration, wherein the hub seal sits in the casing, while the rest of the seals are in the rotor, and where the apex seal has to cope with different wear conditions along its length and perhaps also different lubricating conditions as, under centrifugal force, droplets of oil are carried to one end of the seal.

In my view the other differences cited do not significantly change the balance in favour of the ‘slant axis’ rotary machine, but I hasten to add that this does not detract from the work done by the authors in extending our knowledge and testing the possibilities of this un- explored configuration of the rotary engine.

D. P. Hutchinson Member I see many, and vet a few, of the considerable number of rotary engines that are being offered for consideration at the present time.

Many inventors have not studied their thermodynamics fundamentals adequately and even more have little idea of the basic mechanisms required to make their machines a sound engineering proposition, let alone an economic one. These criticisms cannot justifiably be levelled at this

rotary machine. Ingenuity, backed up by a great deal of careful engineering thought, characterizes this work.

It is perhaps unfortunate that there are so many variations available and but one brief paper to discuss them. The two cycle version selected for detail development appears theoretically ideal as a compression- ignition variant; this is mentioned on page 748. However, the high pressures involved would impose a stringent sealing requirement and there is no further mention of a C.I. version. The choice of this form of the engine further clouds the assessment of its capability since a supply of compressed air is needed to scavenge the engine and it is not clear what contribution this makes to the measured power output of the engine. Including this contribution, the 18 kW obtained (p. 750), or even nearly double this figure (mentioned by J. M. Clarke), is still far short of the 118 kW anticipated (p. 748) and it does seem a pity that development of this interesting form of engine has apparently had to be curtailed before achieving even more convincing output figures. We are also left wondering whether the 4-cycle version might be better, particularly in the spark-ignition version.

In detail I feel that too much emphasis has been placed on low inertia leading to high-speed operation. Surely the latter depends much more on combustion rate than inertia of moving parts and it is not clear how much turbulence can be achieved in the combustion chamber in order to improve it. The low inertia is an advantage when rapid accelerations and decelerations of the engine are required.

This is undoubtedly one of the most interesting and promising rotary piston engine designs so far produced.

A. M. Laws Member My comments fall into three distinct categories-sealing, engine bulk and comparative costs.

Fig. 10 shows the cross-section of a 2 litre unit and this has been used to draw comparisons with other engine types.

The minimal extent of the directly oil-wetted portion of the sealing grid and relatively high frictional losses referred to in the paper of 100 kN/m2 in an indicated mean effective pressure of 400kN/m2 at 1775 rotor rev/min stimulates an interesting comparison:

* Inclusive of valve seats.

The comparison is further illustrated in Fig. 21 which depicts frictional losses taken from dynamometer tests at Associated Engineering Developments Ltd on production reciprocating and Wankel engines, together with the single point result from the paper.

It would seem from its configuration that the National Gas Turbine Establishment engine could readily suffer from an excess of lubricant in the combustion chamber, leading to high emissions and plug fouling, unless stringent oil control measures and maybe capacitor ignition were employed. It would be instructive if the authors could


ROTARY PISTON MACHINE SUITABLE FOR COMPRESSORS, PUMPS AND I.C. ENGINES D263

0 0 2000 3000 4000 rn

ROTATION SPEED rev/min (I) N.G.T.E. engine.

(11) 4-cylinder reciprocating engine. (111) 2-rotor Wankel engine.

Fig. 21. Frictional losses in different engine types

r - - I 1 J IJ I, I

- N.G.T.E. engine.

- - - - 2-rotor Wankel engine. 4-cylinder reciprocating engine.

Fig. 22. Comparison of overall engine size

comment on any lubrication problems encountered, and the palliatives used.

With regard to engine bulk, the proposed type compares favourably with orthodox reciprocating engines of both in-line and ‘vee’ configuration (Fig. 22); indeed, it shares with the second generation of Wankel engines the size advantages so familiar with rotary types.

No details of cost have been given, but in view of the generally high level of finish, abundance of spherical forms and multiplicity of gears, one wonders if it shares with the Wankel the necessity for substantial investment in high precision machine tools; some comment on this important aspect might be appropriate in view of the promising nature of the whole concept.

B. Lawton Graduate In the main I wish to discuss sealing problems for engines of the type developed by the authors.

TRAILING , LEAD1 N G ,TRAIL - . 4

I 2

10 E E . 78 LL

HIGH PRESSURE ROTOR ROTATION degrees SHAFT SPEED 500 rev/min MOTORED

Fig. 23. Computed seal contact force

The authors have made great progress in reducing static leakage area from 20 mm2 to 2 mm2. At the Royal Military College of Science we have measured the leakage area of a 500 cm3 per chamber diesel-Wankel engine and found it to be 1 mm2. However, the achievement of a low static leakage area is one problem; the maintenance of low leakage in a working engine is another. It is convenient to imagine a ‘dynamic’ leakage area superimposed on static leakage area when an engine is operating.

Dynamic leakage is illustrated in Fig. 23. This shows computed seal contact force for one revolution of a rotor. The calculation was made for a motored 500 cm3 per chamber diesel-Wankel engine. At many points in the cycle there is zero contact force and hence zero contact between apex seal and trochoid. Regions of zero contact are regions where dynamic leakage occurs and this is superimposed on static leakage.

It seems to me that the comparatively poor performance, to date, of the authors’ engine can largely be attributed to poor sealing when the engine is in motion. Sealing must be aggravated by relative motion between seal and rotor which is caused by the use of a plane stator seal surface rather than a suitably curved seal surface. Could the authors comment on the feasibility of using a curved stator seal surface. Also, if the authors agree that it is dynamic leakage rather than static leakage that is now the problem, do they have any suggestions for improving the dynamic performance of their seals ? Have chatter marks been observed in the authors’ engine similar to those that are so characteristic of Wankel engines ? Has sufficient running been done to indicate wear rates for the seals ?

Fig. 16 was puzzling at first because the compression pressure seems high. Could the authors give the air supply pressure and also the volumetric efficiency of their engine ?

Finally, in order for us to form an assessment of the combustion chamber could the authors tell us the surface area: volume ratio and the proportion of ‘dead’ volume at the minimumvolume position? Perhaps wecould have these data for both two-stroke and four-stroke arrangements.


D264 DISCUSSION ON J. M. CLARKE, D. F. WALKER AND P. H. HAMILTON

Volumetric&deaCy 0.8 0-9

Tipspe.d(ms-’) 40 80 Dynarmchead loss

Useful head 0.1 0-25

E. Markland Member Anyone interested in this class of machine as an engine will naturally ask what was the net power output and fuel consumption of the experimental unit. Neither of these is given in the paper. The value quoted for measured brake power, 18-1 kW, is quite meaningless-even misleading- in the absence of the corresponding figure for input power to the supercharger which is necessary to run the engine. Will the authors please give us their estimate of this, on the basis of air consumption and supercharge pressure used? May we also have the figure for fuel consumption ?

The authors diagnosed the early failure of the engine as a mixture of poor combustion and poor sealing and have pursued the latter with apparent disregard of the former. However, they now assert that combustion is satisfactory. How can this be reconciled with the statement made, on presentation of the paper, that large quantities of unburnt fuel are discharged in the exhaust ?

Siugle rotor Hooke’s coupling

A. Moore Glasgow and R. A. Meir Glasgow We should like to comment on some of the critical features associated with these machines when they are designed as compressors.

When introduced to these machines by the authors we were impressed with two particular attributes : their high speed capability and their apparent mechanical simplicity. Our work has largely been devoted to analysing and assessing compressor designs which exploit these two features.

If one wishes to run a rotary machine at tip speeds in excess of 50 m s - l then contact seals may be eliminated. Acceptable volumetric efficiencies are obtainable with controlled clearances between rotor and casing. In our analysis we have assumed that the effective leakage gap can be held within 0.1 per cent of the nominal spherical radius. It is essential in the case of a high speed compressor to take account of all the dynamic head losses. In our analysis head losses incurred in both sets of ports along with those associated with the changes in angular momentum given to the gas in passing through the compressor are included.

If one analyses the authors’ design of compressor using the ports suggested in Fig. 9 at the stage pressure ratio 3 : 1, then one finds that the performance is dominated by the dynamic head losses. These losses are attributable in the main to the restricted exhaust port area.

Some of the results of this analysis are given in Table 3. From the results of this work we would suggest that

Table 3

Pressure ratio = 3: 1

_ _

Modified Hooke‘s coupling

Fig. 24

it is unlikely that the overall adiabatic efficiency of such a compressor could exceed 60 per cent.

We are presently engaged in evaluating a modified version of the basic Hooke’s coupling compressor which alleviates the exhaust port restriction. Initial estimates suggest that tip speeds up to 100 m are possible without the overall adiabatic efficiency falling below 70 per cent when the pressure ratio is 3: 1.

Mechanical simplicity is exemplified by the sliding apex machine. Fig. 24 shows a simple blower of this type built at the National Engineering Laboratory. When tested in its crudest form it was found to be noisy. This noise was attributed largely to knocking at the apex associated with the couple C, (equation (23)). Subsequent work involving the use of resilient material to control the preload in the axial direction encourages us to believe that this problem may be overcome.

The sliding contact suggests a tip speed limit of about 10 m s-l . With such a limitation a machine of the simple type shown is not suitable for pressure ratios greater than 1.1.

As the rubbing action at the apex already limits the tip speed, the machine lends itself to sealing. A sealed sliding apex blower appears to be suitable for low capacity duties and pressure ratios up to about 1.5: 1.

J. M. Clarke Member, D. F. Walker Member and P. J. Hamilton Pyestock (Authors) F. Feller suggests that this class of rotary piston machine be referred to as ‘slant axis’ as opposed to ‘parallel axis’ for the Wankel and this seems eminently sensible. It is a descriptive term and unlike ‘precessing’ it is easily enunciated. Other titles such as ‘wobble’ and ‘skew axis’ have been suggested but contain undesirable undertones of loose motion and misalignment. The use of an inventor’s name would be unfair applied to the whole class since similar Hooke’s coupling machines have been seen before. Inertia loads are not an important consideration at low speeds and the parallel axis engine like the reciprocating engine works satisfactorily in spite of them. However, high output versions will run faster and racing reciprocating piston engines certainly fatigue from inertia induced loads. If there was as much freedom to choose combustion chamber geometry in the parallel axis device as there is in



Slant axis rotary piston*

2 4 e I 4-cJrde

the slant axis device perhaps there would be rotaries with more rapid combustion, higher speeds and higher outputs available now.

Cooling of conventional piston crowns is achieved without force-fed cooling oil. In the slant axis rotary it is much simpler to supply oil under pressure for cooling purposes and with proper design of passages it appears possible to cool the rotor adequately in this way.

We would like to thank R. A. Meir and A. Moore for their observations on the possible compressor applications, and wish them success with their researches in that direction. It may be worth remarking that the relative importance of port aerodynamic losses and inertia loads depends on the gas used so that, for instance, the lighter gases such as ammonia and helium permit much higher fluid speeds and therefore offer more scope for using the basically low inertia loads.

A. M. Laws makes some unfavourable comparisons with the parallel axis unit on grounds of ‘oil wetted seal length’ and ‘friction drag’. It is assumed that the circumferential tip seal in the rotor is not directly wetted by oil. Although we have not done so it would be possible to feed oil through the outer casing at a point never crossed by the tip seals and on their low pressure side so that they then become ‘wetted’. The comparison of friction mean effective pressure (m.e.p.) with indicated m.e.p. implies by using the ratio that our problem is a high friction drag. Actually it is our low indicated m.e.p. attributable to leakage which is the trouble. It is also worth noting that our bearing areas are probably greater than necessary since they were originally sized for a compression-ignition duty. In our experience friction drag compares favourably with conventional compression-ignition engine experience (12). A computer simulation of the test conditions corresponding to the results in Fig. 16c gives 778 kN/m2 i.m.e.p. the friction m.e.p. would not rise much, so that point (I) in A. M. Law’s Fig. 21 would probably move below curve (111). Concerning the number of gears it should be stressed that for 4-cycle designs it should be possible to dispense with both the layshaft and bevel shaft and simply use a static bevel pinion. This also achieves a smaller unit. Regarding the size of an investment in machine tools, a transition to any rotary engine involves such an investment. We suspect that differences in investment needs between rotary engines are very much smaller than the basic cost of a change to a new production line.

A. Egberongbe asks about thermal loadings. These would indeed be more severe on the rotor than those of a reciprocating piston. Heat to the rotor cooling oil depends on the extent to which the rotor is cooled by seal contact with the casings and here a long seal length tends to offset a large exposed area. It is difficult to make accurate estimates. We have, however, measured rotor tip temperatures of about 260°C. The quantity of oil needed for lubrication of the seals is very small-perhaps as little as 0.5 per cent of fuel. With oil control of an adequate standard past the hub rings there need be no serious accumulation of oil in the combustion chambers. In the preferred configuration of a 4-stroke spark-ignition engine described below there would not be a combustion chamber in the casing.

B. Lawton raises a number of interesting points concerning cyclic variation of seal performance and loads. The

Reciprocptins PB-t

understanding of dynamic seal behaviour is important to the development of rotary engines whether parallel axis or slant axis. Our experience, however, has been that statically tested leakage areas have given the performance under running conditions which could be estimated by assuming constant leakage area throughout the cycle. We would not expect to achieve measured performance if the seal left the surface for substantial periods as implied by Fig. 23. Surface markings too suggest that apex seals remain in contact, On the other hand our experience with soft seals in a parallel axis machine confirms Fig. 23 to the extent that wear on the apex is greatest at an angle corresponding approximately to B. Lawton’s calculated peak loading. The hard chrome surfaces used on the end casings have tended to crack from thermal strains and poor adhesion but there seems to be no sign of chatter marks. Perhaps this is because there is no apex seal ‘ski-jump’ in the 2-cycle slant axis machine. We have approximated the exact surface by a flat one to avoid manufacturing complications. In terms of extra apex seal movement relative to the rotor this costs about f0.3 mm. The expression for the difference between a flat surface and the actual surface is r{sin x(1 -cos2 O)+sin x sin2 8 cos 2t,h

- cos x sin e( 1 - cos 0) sin 2+ cos t,h} where r is the distance of the seal from the centre, r sin x is the distance of the seal from the central plane of the rotor, and 0 and t,h correspond to the main text. Seal movement is less than this if the seal uses a large radius apex.

Table 4 gives figures for surfaces and volumes and a comparison with a reciprocating unit. The Am,J Vmln ratio is normally closer than this comparison indicates because rotary engines having equal power to their reciprocating equivalents normally use larger chambers.

Regarding the remarks by D. Hutchinson and E. Markland we would not like the results obtained to be

82-33 m-’ 70.0 m-l I / Am*. Vmm - 34-5

0.164 m3

Minimum possible clearance I 0 I o - W m 3 I Lessthan 0 volume


D266 DISCUSSION ON J. M. CLARKE, D. F. WALKER AND P. H. HAMILTON

taken as representing the potential of the machine-far from it we believe. In trying to put our results into context and to explain the high levels of compression pressure observed in Fig. 16 the following points deserve mention.

(1) The basic arrangement was intended for compression-ignition operation. Using spark-ignition so that lower compression pressures and leakage losses were feasible (i) it was not possible to provide an ideal combustion chamber having all its volume in a rotor cutout because the rotor was too narrow to accommodate both combustion chamber and coolant passages, and (ii) the scavenge-blowing arrangement meant that some fuel added in the inlet manifold by-passed the cycle and flowed directly down the exhaust.

(2) Given a further stage of work on this unit it is felt desirable (i) to transfer to the 4-stroke version so that the unit is self-contained, (ii) to place the entire seal grid within the rotor and (iii) to place the whole combustion chamber within the rotor.

(3) The compression pressures are high in Fig. 16c because (i) the supply pressure was 125 kN/m2 with an atmospheric exhaust of 100 kN/m2 and (ii) the pressure when the inlet port finally closes is higher than the supply pressure (1 52 kN/m2) according to our simulator programme, because by this time the chamber volume is being reduced rapidly and flow is reversed through the inlet port.

(4) When the running conditions are repeated on the simulator and there is no seal leakage the indicator diagram appears as in Fig. 25 having an indicated mean effective pressure (i.m.e.p.) of 778 kN/m2, Evidently a t the present leakage levels the i.m.e.p. is approximately half the ideal value.

(5) Using a rig compressor to supply higher air pressure than the exhaust tends to raise the expected shaft output because (i) with the asymmetric port timing pumping work can be done by the charge air on the rotor, (ii) the density of the charge air is increased and (iii) the shaft power output could be debited with the necessary compression work of the blower; on the other hand part of this work is normally supplied by an expander. Taking the results for which a 400 kN/m2 i.m.e.p. is quoted the pumping work m.e.p. is of the order of 2 kN/m2 and the compression work of the blower requires a torque corresponding to 45 kN/m2. The same charge air density has been used for the simulated ideal case as for the measured case. Making these allowances illustrates that the machine is certainly not being blown round by pumping work from its supply air pressure and the claim to have reached about half ideal (no leakage) i.m.e.p. is valid after allowance for blower work. Fuel consumption has to be higher than normal because i.m.e.p. is lower due to leakage and because by-passing of excess scavenge air, not necessarily inadequate combustion, causes unburnt fuel to be discharged into the exhaust.

A. J. S . Baker asks about cyclic loads on timing gears. It is quite true that rotor inertia will tend to oppose fluctuations in mainshaft speed through the timing gears. It seems very difficult to calculate the magnitude of these loads. Such cyclic loads would have to cause gear backlash oscillations if they exceed the relatively small steady torque caused by the difference between rotor bearing drag and seal drag. We would expect to detect chattering of the gears in this event but we did not. Except for instances

Fig. 25

where cold backlash was too small there were no mechanical troubles attributable to the gears. All gears incidentally were milled. The rotor can be made thicker at the design stage simply by siting the seals further from the centre plane of the rotor. In the case of designs having all the seals in the rotor, the combustion chamber can be situated partly in the hub sphere of the rotor. In this way low compression ratios could be accommodated without using thicker rotors. Our calculations have shown volume variations to be sinusoidal to within f0.06 per cent of the sine semi-amplitude. We cannot agree entirely with the suggestion that the turbulence may be low in the in- coming air since the individual chambers will see a cyclic variation in mass flow similar to that in the conventional engine. Air motions within the chambers must affect heat fluxes and probably differ between engines much more significantly than the surface areas. It seems certain that in this slant axis engine as in the parallel axis engine heat losses are higher than those of the reciprocating engine but the effect on fuel consumption may be partly offset by lower friction. The hub rings have in fact been designed to recirculate oil using the oil scraper and tapered land technique but a poor surface finish of both hub and grooves did lead to excessive oil passing this arrangement. Pressures were in fact measured between the end seals. Any detectable rise (3 kN/m2) presaged failure of the hub compression ring. The capstan effect is relevant to the hub and tip seals and the proper positioning of their anchor points relative to the rotation direction. In fact the hub ring was anchoredso that three-quarters ofthe circumference wrapped on while a quarter wrapped off. The tip seals were located so that their half-circumference lengths wrapped off. This latter choice was based on discussion



Low external inertia forces

compression ratio Potentially high

rather than experience of worse arrangements, but the hub ring anchor point followed less satisfactory experience using full circumference wrap off. The late R. C. Cross was in fact responsible for the supply of the steel hub rings, and they never broke.

Since the paper was written higher powers were obtained, the highest being 30 kW at a rotor speed of 2325 rev/min corresponding to a b.m.e.p. of 380 kN/m2. At the inlet air conditions used for this particular test the air blower would absorb about 133 kN/m2, but scavenge was excessive and exhaust back pressuring could supply a proportion of this by allowing expansion work.

Table 5 represents an attempt to summarize our view of the position of the slant axis rotary engine. Any rotary having an attraction commercially should offer over the reciprocating engine a smaller size and the absence of valve gear. To succeed in performance it must use sliding contact seals which precludes numerous external-axis machines such as the Tri-Dyne. It should have low inertia induced external forces for comfort and silence and low inertia induced internal forces for low stresses at speed and to allow potential for development to high speeds. For possible development to compression ignition and to allow options in combustion chamber shape it should have potential for high compression ratio.

The kind opening remarks by many contributors to the

- 4 J I J J

J d J / - J

-----

Table 5

Compactness I - I d 1 4 I d

No valves I - 1 4 4 4 I d ~

Sliding contact seals

Low internal inertia forces

discussion are acknowledged with gratitude. Having reached this stage it is our earnest hope that the potential of the slant axis rotary piston machines will be explored much further.

R E F E R E N C E

(12) MILLINGTON, B. W. and HARTLES, E. R. ‘Frictional losses in diesel engines’, S.A.E. Paper No. 680590, September 1968.

Proc lnstn Mech Engrs 1972 Vol 186 6211 2

Documents

A new class of rotary piston machine suitable for compressors, pumps and internal combustion engines