The Mechanisms of Piston Ring Operation

This article of the attempts to condense the available data on the various aspects of piston ring performance into a more complete and readable form than that in which it is otherwise available. In those areas where the information is insufficient, or where the operating conditions may vary substantially from one application to another, the descriptions of ring operation are essentially rather tentative.

In the subsequent final section of the report an attempt has been made to list some of the more positive conclusions from the research, in a form which is convenient for direct reference for design.

Sealing and Blow-By

This section deals with all aspects of piston ring sealing including blow-by.

The Mechanism of Sealing

A piston ring seals a pressure above it by contacting the cylinder bore with its outer face and the piston groove at its lower side face. Contact with the cylinder bore is effected initially by the outwards expansion of the elastically tensioned ring; and it is then the pressure drop past the ring which tends to force it against the lower side of its groove.

Once a seal is made at the lower side face, the full pressure above the ring is communicated behind it via the upper side clearance, and it is then the pressure itself, acting above and behind the ring, which provides the major sealing force.

At least in the region of small separation between the surfaces, oil can be assumed to be present between them either as a continuous film, or in the interstices between them if they contact at high points. Disregarding any effects of ring motion, the pressure in the oil films will be hydrostatic only, and will fall progressively across the sealing surfaces from the high gas pressure on one side, to the low gas pressure on the other.

The situation is illustrated in Fig.4. The hydrostatic pressure distribution opposes the direct gas loading so that the net loading due to the pressure drop is reduced. For parallel sealing surfaces, the hydrostatic pressure distribution will be linear, and the net loading due to gas pressure will then be half the pressure drop.

For the more general case of non-parallel surfaces, the region of reduced separation will constitute a main sealing band at which most of the pressure drop will take place (see Fig.4). A degree of pressure balancing will be provided on the high pressure side of this main sealing band, and the net gas loading will be some fraction (between 0 and 1) of the pressure drop depending on where, across the sealing faces, the main sealing effect is obtained.

Fig 4 Effective Ring Loading Due To Gas Pressure

  • Net Radial Gas Load = Area A – Area B
  • Area C = x w Δp
  • Effective Pressure due to Net Axial Gas Load = y Δp
  • Effective Pressure due to Net Radial Gas Load = x Δp

Thus, the fraction (for radial loading) is about 0.5 for parallel or centrally crowned rings, about 0.9 for up scraping rings, and about 0.1 for down scraping rings. Similarly, for axial loading, the fraction depends on the contact geometry on the side face.

Since the dominant loading on a piston ring is that due to gas pressure, it is obviously important to determine the factors which affect the pressure distribution in a ring pack, and to be able to make at least a rough estimate of the pressure distribution which can be expected in any given practical situation. This is considered in more details in the following section.

Pressure distribution in a ring pack

The pressure distribution in a ring pack is determined by the way in which the rings and clearance spaces of the pack restrict the flow of gas leaking through the pack from the cylinder. For correctly functioning rings, under most operating conditions, gas leakage past the running face and the sealing side face of a ring is negligible compared with the leakage through the ring gap. For rings which are unchamfered at their lower outer edges, the effective leakage area of the ring is then the product of the ring gap width and the clearance of its supporting land from the liner.

During early run-in or change of engine load, or when running with poorly fitting rings or an inadequate oil film on the liner, blow-by past the sealing surfaces can occur, resulting in the effective leakage area being greater. For the purposes of determining the pressure distribution, however, normal operation must be assumed.

The restriction presented to leakage flow by the clearance spaces between the piston lands and liner, and by the side and back clearances between the ring and its groove may normally by neglected when compared with the restrictions presented by the ring gaps to leakage flow.

Thus, for the purpose of determining pressures, the ring pack may be viewed as a series of orifices corresponding to the ring gaps, separated by chambers corresponding to the spaces between the rings.

The usual way in which the pressure is distributed through the pack, and how it varies during the engine cycle, can then be determined by the application of gas dynamics theory. Good agreement has been found between this theory and practical measurements (25), and computer programmes are available (52) for determining the pressure distribution and its variations.

The results obtained from calculations of this type are broadly as follows:-

  1. In the case of a steady cylinder pressure above the pack and identical rings, it is found that once a steady distribution is established most of the pressure drop occurs across the bottom ring. This arises because although the mass flow through each ring gap is the same, the mean density of the gas flowing through the bottom ring is lower, and consequently the volume flow is higher, resulting in a higher pressures drop. This situation arises in practice in certain types of low speed multi- stage compressor.
  2. In the case of the cylinder pressure taking the form of a pulse of very short duration, nearly all the pressure drop occurs across the top ring since there is insufficient time for the space between the top and second rings to be charged to any significant pressure.

In practice, the normal situation lies between these two extremes, and the extent to which the pressure penetrates the pack is determined broadly by the time which would be taken for the pressure below each ring to rise to the same pressures as that above it (as determined by the ring leakage area and the volume of the space below it) compared with the time that the pressure pulse exists above the ring.

This concept of relative filling time can give a rough indication of the type of pressure distribution to be expected (3). Fig.5 shows how the pressures in the cylinder and in the spaces between the rings, vary through the cycle for the case rings appreciably greater than 1); and (b) a slow running engine with larger ring gaps (relative filling time of rings about 0.4).

The significance for ring operation of these different pressure distribution situations and the resulting ring movements, will be discussed in the next section, but before proceeding to this it is useful to note in more detail the factors which affect the relative filling time, and to estimate values of the relative filling time which can be expected in practice. From the estimated relative filling time in any given case, and reference to Fig.5, it should then be possible to obtain a reasonable idea of the type of pressure distribution to be expected. Alternatively, more accurate direct calculations using a computer might be used.

The relative filling time (3) is given by:

T = V / (a k Cto)


  • V = the volume of the space between the ring under consideration and the one below it.
  • a = the effective leakage area of the ring.
  • to = the duration of the cycle of high pressure above the ring, which may be taken as a quarter of an engine revolution (3).
  • co = the maximum velocity of the gas flow through the ring leakage area, and may be taken as the velocity of sound corresponding to the gas temperature, that is
  • k = a factor for correcting co to obtain the mean velocity during filling of the space below the ring, and may be taken as 0.5

figure 5 multiple piston rings

(B) Short Relative Filling Time (≈ 0.4) (Slow running engine with poorly sealing rings)

Fig 5 Pressure Distribution in Two Ring Packs

Taking k as 0.5 and to as one quarter of an engine revolution, the relative filling time is given as:

T = 8 x Vn / aCo or t = 0..11 x VN / aCo


  • n is engine speed in rps N is engine speed in rpm
  • V is in cm3 V is in in3
  • A is in mm3 a is in in2
  • Co is in m/s Co is in ft/s

The temperature of the leakage gas and hence co are somewhat uncertain, but for the purpose of the rough estimation which follows, a figure of 600 m/s (2000 ft/s) corresponding to a typical gas temperature, will be used, giving:-

T = 0.013 V n / a or T = 5.5 x 10-6 V N / a

  • n is in rps N is in rpm
  • V is in cm3 V is in in3
  • A is in mm2 a is in in2
  • T = 0.013 V n or T = 5.5 x 10-6 V N

Thus higher speeds, tighter rings, and larger clearance spaces in the pack all tend to raise the filling time. For the purpose of calculating T from this expression, the ratio V/a must be estimated. For unworn, clean piston assemblies and for square-section, un-chamfered butt-jointed rings, lands and grooves, typical dimensions taken from (24, 58) and the corresponding values of V and a are indicated in Fig.6. It can be seen that the ratio V/a varies from about 20 to 80 m over the engine size range. Using these values in the above expression, T varies over the engine speed range as indicated in Fig.7.

Considerable variations from these values are possible, however, not only with different assembly designs, but also during the operation of an engine. Effects of design include variations from the ‘typical’ dimensions shown in Fig.6 and the use of different ring types and different ring gap designs. For example, the use of twist rings with a large cut out from the ring section, will greatly increase V and the use of Duolap joints will appreciably reduce a.

Effects of operation include: variation of the land clearance at the ring joint due to lateral piston movement and ring rotation; variation of bore diameter up and down the cylinder; dimensional changes due to thermal expansion and the changes of temperature distribution with varying engine load; increased leakage area due to poorly fitting rings; and, over a longer time scale, variation of clearances with wear and deposition of carbonaceous residues.

The latter effect will normally reduce the relative filling time by at least a factor of ten by the time the rings are due for replacement, and a similar reduction if likely with poorly fitting rings. The effect of such a reduction on the relative filling time is indicated in Fig.7. In spite of these various effects, Fig.7 may be used for initial guidance on the value of the relative filling time, and a much more accurate estimate is normally possible from a detailed assessment of the values of V and a in any individual case. The type of pressure distribution to be expected can then be roughly estimated, and with this knowledge it is then possible to estimate cycles of loading and movement for the rings, as indicated in the next section.

Fig 6 Typical Compression Ring and Groove Dimensions and the Corresponding Ring Leakage Areas and Inter-Ring Volumes.

Speed, n (rps)

Fig 7 Typical Variation of Some Ring and Engine Design Parameters With Engine Speed.

Cycles of ring loading and movement

In general a piston ring is subject to the following forces:

  • Radial and axial gas loading
  • Elastic ‘tension’
  • Inertia
  • Running and side face friction
  • Running and side face loading

The running and side face loadings are the ones which determine the contact conditions at the bearing surfaces and which, when zero, imply that the bearing surfaces may separate. These face loadings depend on the balance of forces in the radial and axial directions respectively, and cycles of ring loadings can therefore be investigated by considering how these other forces vary in magnitude and direction during an engine cycle.

Before discussing particular situations which may arise, it is useful to have a general idea of the magnitudes and characteristics of the various components of ring loading, and these are summarised in Table 1 overleaf. A number of the ‘typical values’ shown in the table are related to engine speed on the basis of the typical dimensions of an engine of a given sped, and for convenience these have been plotted on Fig.7. Whilst deviations from these values by a factor of 2 either way are likely in practice, the values nevertheless give a useful general idea of the size of the various loading components, together with their current trends over the practical range of engine type.

It is apparent from Fig.7 that the loading situation changes considerably across the range of engine speeds. To illustrate the range of effects which can occur, the cycles of ring loading and movement for, firstly, large slow-speed engines, and then smaller, higher-speed engines, are considered in turn below.

Slow Speed Engines

By reference to Fig.7 it can be seen that both the inertia and elastic pressures for the slow speed engines are small, and the ring movement and loading are consequently determined very largely by the pressure distribution and its cyclic variation. Also, the relative filling time tends to be short, and the pressure distribution of Fig. 5(b) is typical for engine of this type. Englisch has described the behaviour and operation of rings for this type of engine in great details in (25).


It can also be seen from Fig.5 (b) that the pressure drop across the top ring actually reverses quite early during the power stroke, and since the other axial forces on the ring, due to friction and inertia, are relatively small, the axial loading on the top ring will reverse at about the same moment. This will cause the top ring to lift from the lower side of its groove and re-seat on the opposite upper side.

This movement, the existence of which is well proven in practice will have several consequences, the more important of which appear to be as follows:

  1. An impact will occur between the ring and the upper side of its groove. Although the associated stresses and wear are not normally severe in themselves, the resulting ring marking is a useful witness to the movement and to the shape of the ring grooves during operation. The stresses may become appreciable and accelerate ring breakage when large side clearances develop due to wear, or if the load reversal is very rapid.
  2. The movement of the ring may well have an important effect on oil distribution and ring lubrication. Firstly, the movement will pump oil out on to the bore from any stores on the upper face of the ring. This may be beneficial in terms of oil distribution to the bore and lubrication of the rings, or undesirable in terms of increased oil consumption, depending on the magnitude of the effect and other factors. At the same time, if the ring has scraped oil up on to its lower edge at this moment in the cycle, as is likely, this oil will be drawn into its lower side clearance space and replenish the lower side oil film. This is probably highly desirable as a means of reducing the friction and wear at the side face, and also important as another stage in the circulation of oil round the back of the ring.
  3. The movement may be important as a mechanism of compacting or breaking up groove deposits.

As the cycle progresses beyond the point of top ring lifting, the second and perhaps subsequent rings may lift in the same way as illustrated in Fig.5 (b). At this stage the pack is sealed at both ends, and relatively high pressures can become trapped within the pack, the highest usually lying behind the middle ring(s).

In four-stroke engines the presence of such trapped pressures is unimportant; the pressures fall more or less rapidly as the trapped gases leak from each end of the pack, and are normally fully relieved by the start of the next cycle.

With two-stroke engines, however, these trapped pressures may cause undesirable effects as the rings pass over the upper edges of the ports. When thebottom ring passes into the ported area it rests on its lower side with high pressure behind and above it. As its top edge uncovers the ports the pressure behind cannot be relieved instantaneously, and will tend to jolt the ring out into the ports. These trapped pressures may be up to 15 atmospheres (25) and particularly when a ring horn is unfavourably placed relative to the ports, the resulting bending stresses can be very high and are probably a major factor in ring breakage. Each successive ring will experience this effect, but it will be most severe for the middle rings due to their higher back pressure.

Also, it is possible that the upper rings, and especially the top ring, may still be seated on their upper sides when the pressure below them is relieved into the ports. If this happens the affected ring will have to jump back across its groove and, due to the very rapid reversal of the pressure drop, this jump may be violent and contribute appreciably to lower groove-side damage.

From this discussion it can be seen that in all slow-speed engines the upper rings at any rate lift at least once in their cycle. In four-stroke engines all the rings will be lifted again when the piston passes through TDC exhaust and descends on the suction stroke. In two-strokes it is possible, but unlikely, that all the rings may lift again when crossing the port bars, due to the friction force outweighing the ring inertia, but certainly the lower rings of the pack can be expected to lift up in the region of TDC, due to their inertia, before appreciable gas pressure penetrates to them.

There would appear to be one further loading situation which may have an important affect on ring loading, and which is more pronounced on low speed engines. This concerns the effect of side face friction on the running face loading of the top ring when it is carrying high pressure loading near TDC. At this stage in the engine cycle, for low speed engines at any rate, the elastic and inertia pressures will be insignificant. The running face loading will therefore be determined by the radial gas loading and the friction force on the side face. If the side face friction coefficient is large for any reason, and the ring has a high aspect ratio (radial thickness to axial width) the change in effective radial pressure due to the side friction may be an appreciable proportion (up to a maximum of probably about 0.25) of the pressure drop across the ring. If, at the same time, the net radial gas loading is a relatively small proportion of the pressure drop because the running face contacts the liner towards its lower edge, it is possible that the net radial gas loading could be less than the side friction. In this situation any movement of the piston away from the bore will cause the ring to be dragged away from the liner, resulting in blow-by past the running face, and a risk of subsequent scuffing.

One practical situation in which this could arise is with taper faced down-scraping rings in a cylinder which opens out upwards in the upper part of the bore. It could also arise on the anti-thrust side of a cylinder when the piston crown moves across the bore around TDC, especially if the crown moves first so that the piston tilts in the direction which encouragers lower edge contact on the ring running face. In practice this type of piston movement may often occur in trunk piston engines due to the small end bearing torque. A further effect in such cases is that the ring loading on the thrust side of the cylinder will be greatly increased, not only due to the additive effect of the friction force on this side, but also due to the piston tilt, which encourages upper edge contact on the ring running face so that the radial gas load is a high proportion of the pressure drop.

Although it is unlikely that the extreme conditions leading to a ring lifting away from the bore are frequently met, it is probable that all trunk piston engines experience some degree of loading asymmetry at TDC due to these effects, and this is the most likely explanation of the greater wear step which is nearly always found on the thrust side of worn liners from these engines.

As mentioned previously, the larger slow speed engines are probably more affected in this way. This is partly because their ring aspect ratios tend to be larger, but also because in the high speed engines, firstly side friction is reduced over TDC due to a measure of axial loading relief from ring inertia; and secondly, higher elastic pressures are used which help to maintain radial loading.

Higher Speed Engines

Compared with the slow speed engines just discussed, higher speed engines are characterised by longer relative filling times and higher inertia pressures, and normally use rings which have higher elastic pressures and greater heights in relation to their widths (see Fig.7)

As a consequence of the longer relative filling time, the pressure distribution is typically as shown in Fig.5 (b). By comparison with Fig.5 (a), it can be seen that one main difference with the longer relative filling time is that the top ring carries nearly all the gas pressure drop across itself alone, so that its cycle of loading is severe, whilst that of the lower rings is very light. This accounts for the use of a smaller number of rings in the higher speed engines; in many cases any rings in addition to the top ring would appear to have a standby role only.

The second major consequence of the longer relative filling time is that the timing of the pressure reversal is retarded to appreciably later in the power stroke. This will effect the timing and incidence of ring lifting.

For two-strokes, the pressure reversal may not occur at all if the relative filling time is long enough, due to venting of the spaces between the rings at the ports before they reach sufficient pressure. Furthermore, for engines with typical ring characteristics, and running face friction coefficient, the inertia pressure is likely to exceed the friction force by the time the rings have entered the port belt, and as the stroke progresses, the balance will be increasingly downwards.

Consequently, the rings will probably remain on their lower sides down to BDC. Beyond BDC, at least up to the point of acceleration reversal about 80° before TDC, all the forces on the rings are downwards so that lower side contact will persist up to this point. Whether any of the rings lift beyond this point will depend on how quickly the axial gas pressure and friction loading build up on each ring compared with the inertia force.

For most engines running at their normal speed and full power, the cylinder pressure will rise fast enough on compression to hold the top ring down.

On the other hand, the pressure below the top ring (provided this is sealing reasonably effectively) will not normally rise nearly fast enough to prevent all the lower rings jumping up across their grooves. Where they will normally remain until well down the power stroke.

It appears, then, that for two-strokes which have long relative filling times, the top ring may not lift at all during normal full power running. This will reduce oil circulation around the back of the top ring, and as a result probably adversely affect the upper cylinder lubrication and increase the tendency to ring sticking. It is also possible that the lubrication of the lower side of the top ring could be impaired.

For four-stroke engines, even when the relative filling time is very long, a pressure reversal across the top ring should occur shortly after the exhaust valve opens, and this may result in the ring lifting briefly to release the pressure stores between it and the second ring if the inertia force is not too great. Even if the top ring does not lift at this stage it will certainly do so, in company with the other rings, in the region of TDC at the end of the exhaust stroke. The lower rings will probably also lift again when approaching TDC on compression. At this stage in the cycle, in the same way as has been described for two-strokes, the top ring should normally be held down by the rising cylinder pressure.

In contrast to this normal behaviour, it is possible in some circumstances, in both two and four strokes, for the pressure drop across the top ring to fail to rise fast enough during compression to keep the top ring on its lower side, and this can have very important consequences for ring operation and performance. This situation was investigated experimentally by Dykes some 25 years ago (19) enabling him to give a particularly clear picture of ring movement in a high speed four-stroke petrol engine. Up to a certain speed the ring behaviour was ‘normal’ that is, the ring only lifted once in the cycle, approaching TDC on the exhaust stroke, and fell again during the suction stoke. Above this speed the ring also lifted approaching TDC on compression, and dropped back again after TDC on the power stroke.

Besides this additional axial movement, however, Dykes also showed that the ring collapsed radially upon reaching the top side of its groove. The explanation of this behaviour is, firstly, that the increase of speed above the ‘critical’ raised the inertia pressure sufficiently for it to overcome the gas loading as the piston approached TDC. The ring consequently jumped across its groove and made a seal at its upper side. The pressure below and behind it was relieved, and the ring was then collapsed inwards by the higher gas loading on its running face. Blow-by then occurred past the top ring running face, caused the second ring to collapse inwards in a similar manner to the top ring, and then became severs, only ceasing when the piston acceleration reduced sufficiently after TDC to allow one at least of the rings to drop back to its lower side and reseal in the normal way.

It is essential that this type of behaviour is avoided in all normal operation of any engine, and it is therefore important to note, firstly, the conditions in which it is encouraged the initial lifting of the ring are:

  • High speed
  • Low bmep or throttle opening
  • Axially wide rings
  • Poorly fitting rings

The axial lift is probably not very important in itself, and is in any case almost impossible to avoid on overrun. It is the radial collapse which may follow, which must be avoided.

This can be done with rectangular section rings by ensuring that the ring elastic pressure is large enough. Based on a simple analysis of the ring equilibrium in the lifted position, the minimum elastic pressure required is probably a little less than the maximum inertia pressure, and can therefore be calculated quite readily in any given case. Design values should be appreciably greater than this minimum to allow for tension loss during operation.

By reference to Fig.7 it can be seen that the typical values used for the ring elastic pressure and the ring inertia pressure (based on typical ring widths) give a reducing margin of safety against ring radial collapse with increasing speed. This ties up with the general experience of severe blow-by problems in high speed engines. It would therefore seem desirable to use axially thinner rings of higher elastic pressure in higher speed engines than is current practice, and the recommendations of (24), shown in Fig.7, appear to meet the requirement well.

For very high speed operation, and to ensure protection against radial collapse even after a lot of wear, an L-shaped ring proposed by Dykes appears to be a good solution. With this ring, axial ring lift is limited by the horizontal part of the L-shaped section, and extra clearance is left at the top pf the ring to ensure that the high pressure can still get behind the ring to push it out against the bore. It seems likely that this ring could find useful application in many high speed engines, especially where a deeper running face is required for best wear performance.

A final effect to which high speed engines may be particularly, but not exclusively susceptible under some circumstances is retarded pressure build up in the back clearance space behind the top rings due to appreciable throttling of the leakage flow through the ring side clearance. If this occurs it is possible, particularly with ‘radially pressure balanced’ rings, for the running face gas loading to temporarily exceed the outward loading, as the cylinder pressure rises rapidly towards the firing peak. This will cause the ring to collapse inwards, thereby permitting blow-by and throwing the sealing duty on to the second ring.

The delay in the pressure build up behind the top ring depends on the relative filling time of the back clearance space which, for diesel engines, varies from about 0.002 @ 2 rps (120 rpm) to 0.02 @ 130 rps (8000 rpm) for typical ring and groove dimensions.

A relative filling time of 0.02 is likely to cause an appreciable lag of the back pressure, at least in diesel engines where the rate of cylinder pressure rise is particularly rapid, and suggests that, for high speed diesel engines, this effect could also be a possible cause of top ring radial collapse during the compression stroke. For petrol engines the rate of cylinder pressure rise is less rapid, but the relative filling time for the back clearance space tends to be considerably greater (by a factor of typically 3) due to the employment of larger back clearances and smaller side clearances than in the diesel engines. Consequently petrol engines are probably at least as susceptible to this problem as diesel engines. Means for avoiding this problem are, again, a reduction of ring height and an increase of ring elastic pressure, and it will also be advantageous to keep the ratio of back to side clearances as low as possible. Groove packing may also increase the likelihood of this situation arising if the side clearance is most rapidly affected. And poor ring performance could result due to this effect, even with all rings free.

Blow-by can also occur if the cylinder bore is not circular. The effect of deviations in bore circularity has been studied in (21). The permissible deviation, beyond which blow-by may increase sharply, varies with the number of lobes in the deviation, and with the ring gap position. In the tests of (21), for a purely oval (2-lobed) deviation, the permissible out-of- roundness (difference between biggest and smallest diameter) on a 240 mm diameter bore was 0.6 mm with a straight cut ring gap lying on a minor axis of the oval. The permissible out-of-roundness was more than twice as great with the gap on the major axis, and for a 4-lobed deviation, the limiting difference of diameters was reduced to 0.15 mm.

Thus, in badly worn or distorted liners, high tension rings are again desirable to give reasonable conformability of the rings to the bore.