In this article we will discuss about:- 1. Meaning of Compressor 2. Classification of Compressors 3. Performance Characteristics 4. Limitations.
Meaning of Compressor:
Compressor is a device which compresses air/gases or vapours from low pressure to high pressure. It needs external energy input in the form of work. Out of the total work input to the compressor, some work is utilised to compress the fluid while the remaining is lost in overcoming friction, some work is lost to cooling medium, etc.
Compressors are mainly grouped into two categories:
(i) Positive displacement compressor and
(ii) Dynamic compressors.
In a positive displacement compressor, the pressure of the gas is increased by reduction of its volume i.e., by positive displacement of the gas to the delivery side.
While in a dynamic compressor, the kinetic energy imparted to the gas by the rotation of the rotor (impeller) is converted into pressure energy partly in the rotor and the remaining in the diffuser. Thus, the pressure rise is developed due to the dynamic action of gas.
Compressors are extensively used for a variety of applications.
Compressors are classified on the basis of several criteria as follows:
(a) On the basis of design and principle of operation – Based on the design and principle of operation, compressors are classified into two main groups and further sub-divided into sub-groups.
(b) According to delivery pressure – Low pressure (upto 10 bar), medium pressure (10 – 80 bar), high pressure (80 – 1000 bar).
Hyper-compressors are multistage reciprocating compressors with delivery pressure upto 1000 bar.
(c) According to the pressure ratio (as per ASME code) – Fans — Pressure ratio upto 1.1 Blowers — Pressure ratio between 1.1 to 2.3
Compressor — Pressure ratio above 2.3
(d) According to free air delivered (capacity). Small capacity — Upto 9 m3/min Medium capacity — Between 9 – 3000 m3/min
Large capacity — Above 3000 m3/min
The present article deals with Rotary positive displacement compressors and dynamic compressors.
When a flowing fluid having same velocity is brought to rest, it is said to have attained a stagnation state. The final stagnation state is governed by the manner in which it is attained. A reversible adiabatic or isentropic process has a great significance.
For an isentropic stagnation process, die steady flow energy equation simplifies to-
Where ho is stagnation enthalpy and h is its initial enthalpy when fluid is flowing with a velocity of V m/s.
The properties of the fluid at the stagnation state are referred as stagnation properties e.g., stagnation pressure, stagnation temperature, stagnation density, etc. The stagnation state and stagnation properties are denoted by lower suffix o.
During the stagnation process, the kinetic energy of the fluid is converted to enthalpy, thereby pressure and temperature of the fluid increases as shown below-
An isentropic stagnation state is the state, fluid attains when the stagnation is reversible adiabatic or isentropic, shown by process a – b on the h-s diagram. Actual irreversible process with friction or heat transfer is shown by process a – c. It may be noted that stagnation enthalpy hb and hc is same in both the processes. However actual stagnation pressure pc is lower than isentropic stagnation process pb, due to entropy increase in the actual process on account of friction. Now, for an ideal gas undergoing isentropic stagnation process,
For better understanding of the velocity triangles of rotary compressors, the concept of absolute velocity and relative velocity is necessary.
Absolute Velocity and Relative Velocity:
1. Absolute Velocity:
It is defined as the velocity of a moving object as sensed by a stationary observer. In a true sense, no observer can be stationary on the earth as the earth is constantly moving slowly. However, its slow motion can be ignored.
2. Relative Velocity:
It is defined as the velocity of a moving object as sensed by an observer who is moving with its own velocity.
To illustrate this concept further, consider an object moving with a velocity V m/s. When a stationary observer looks at this object, the observer can get a correct feel about the magnitude and direction of the moving object. However, if the observer also has its own velocity, V m/s, then the observer gets only an apparent feel about the magnitude and direction of the moving object. This apparent feel of velocity is the relative velocity. Let us consider the following examples-
Here, the relative velocity Vr is the vector difference of the two absolute velocity vectors. Figure 16.2 (b) shows the two moving objects moving in opposite direction. The relative velocity Vr is the vector sum of the two absolute velocity vectors. In general, the procedure to obtain the relative velocity between the two objects moving in their own directions can be stated as follows-
From a common starting point, set both absolute velocity vectors in their magnitude and direction. The line joining the ends of the two absolute velocity vectors represents the relative velocity which is the vector difference of the known absolute velocity vectors, see Fig. 16.2 (c).
Similarly, when a relative velocity and one of the absolute velocities are known, the unknown absolute velocity is given by the vector sum of the known relative velocity and known absolute velocity. The graphical procedure is as follows-
Draw the known relative velocity vector to some scale, in its magnitude and direction. In succession i.e., starting from the end of the relative velocity vector draw to same scale, the known absolute velocity vector in its magnitude and direction. Then the line joining the starting point to end point, representing vector sum, gives the unknown absolute velocity.
Performance Characteristics of Compressors:
Performance Characteristics of Centrifugal Flow Compressor:
The performance of centrifugal compressors is specified by illustrating the variation of delivery pressure and temperature with mass flow rate for various speeds. However, these characteristics are further influenced by other variables like inlet pressure and temperature.
To understand the interdependence of these variables, dimensionless parameters are often useful, such as-
Points of low mass flow at the extremities of constant speed curves can be joined to obtain surge line. While the curve obtained by joining extremities on the right hand side of constant speed curves represents choking limit occurring at low pressure ratio. The compressor is operated only between these extreme limits.
From Figs 16.28 and 16.29, following deductions can be made:
1. The curves are fairly flat at lower speeds and are limited by surging. At high speed, the range is limited by surging at one end and choking at the other end.
2. At a given speed, mass flow rate reduces as the pressure ratio increases.
3. At a given pressure ratio, increase in the speed increases the flow rate with considerable reduction in the efficiency.
Performance Characteristics of Axial Flow Compressors:
It is seen that, for a given value of N/√To1, these curves cover a much narrow range of mass flow rate as compared with those for centrifugal compressors. Also at higher rotational speeds, the curves become steep, nearly vertical. Therefore, the range of stable operation of axial flow compressor is much narrower. Hence, great care is necessary to match the components of a gas turbine plant for avoiding instability of operation. The phenomena of surging and stalling can hardly be distinguished as occurrence of one may lead to occurrence of the other. Stalling of these compressors results in severe vibrations of blade.
Limitations of Compressors:
In the operation of centrifugal and axial flow compressor, there is an instability known as surging. It is caused due to unsteady, periodic and reversal of flow through the compressor when it is operated at less mass flow rate than that corresponding to maximum pressure.
Surging may lead to mechanical failure. The rotor is subjected to alternating stresses during this irregular working and may result in damage to the bearings, rotor blades and seals. In the extreme case, the rotor shaft may bend.
Consider the pressure ratio (head) versus mass flow rate curve of a centrifugal or axial flow compressor as shown in Fig 16.25. The curve consists of part AB having a positive slope and part BC having a negative slope. Point A represents fully closed discharge valve while B represents full open discharge valve.
At point A, the flow rate is zero and pressure developed is called shut-off head. And at point B, pressure head developed is zero when the mass flow rate is maximum. Thus at point B, efficiency is zero.
When the discharge valve is opened gradually from its fully closed condition, the discharge of fluid commences and static pressure gradually built-up due diffuser adding to the pressure rise. Further opening of discharge valve increases the pressure till point B is reached. At this point, efficiency is maximum for a given speed, inlet pressure and temperature.
Now, when delivery valve is opened beyond point B, the mass flow rate increases but pressure rise and efficiency decreases. This trend continues till point C when the pressure ratio approaches unity while the mass flow rate is maximum but the efficiency is zero.
The opening and closing of a discharge valve acts as an external load on the compressor. The compressor operates in response to the external load. The intersection of the compressor curve and the load curve represents the operating point, say as at point D in Fig. 16.25.
When the discharge valve is closed further thus increasing the external load, the back pressure in the discharge line increases. As a result, new operating point shifts to E. The new operating point is possible and stable as the compressor develops greater pressure head to offset the increased back pressure in the discharge line.
With further closing of the discharge valve, external load increases, the operating point shifts in the region A-B. During this operation, mass flow rate is less than the estimated value, say corresponding to point B. The compressor develops less head than that existing in the discharge line.
As such, no discharge of fluid is possible. This leads to momentary flow reversal. A short while after, the fluid from the discharge line leaves and the back pressure would reduce. The delivery of fluid from the compressor would restart and the cycle would repeat with instability of operation again.
Therefore, when the flow rate from the compressor is less than the estimated value, a surge or pulsation commences. Air surges back and forth through the whole compressor rather than giving a un-indirection steady stream of flow. The unstable operation of the compressor prevails in the region of positive slope of the curve shown in Fig. 16.25.
Various methods of avoiding and correcting surging conditions are listed below:
It can be controlled:
(a) By speed control.
(b) By throttling inlet.
(c) By provision of flow control system.
i. By venting discharge to the atmosphere.
ii. By passing fluid from discharge line to suction line.
iii. By providing integral expander on the by-pass line.
(d) By cut-off capacity control.
(e) By inlet guide vanes or adjustable diffuser vanes.
Referring to Fig. 16.25, when the mass flow rate is increased beyond B, the pressure ratio decreases and efficiency falls, as the air angle differs significantly from the vane angle causing the break-away of the air-stream.
This continues till point C where the pressure ratio becomes unity and efficiency zero. Entire power input is lost in overcoming internal friction. The point D on the curve BC, represents the maximum mass flow rate, known as ‘choking mass flow’. The phenomenon of choking limits the maximum mass flow rate.
As pressure ratio falls to unity, the theoretical mass flow rate is maximum. This occurs when the Mach number corresponding to inlet relative velocity becomes sonic.
Stalling of a stage of axial flow compressor is defined as the aerodynamic stall or breakaway of the flow stream from the suction side of the blade aero-foil. It may be due to less mass flow rate than the designed value or due to non-uniformity in the blade profile.
The phenomena of stall precede surging. A multistage compressor may operate in the stable unsurged region with some of its stages stalling. Thus, stalling is a local phenomena while surging is a complete system phenomena. The phenomena of stalling have been extensively researched by Smith and Fletcher.
The non-uniformity of the flow or geometry of blade profile causes blade B to stall. The air now flows on to the blade A at an increased incidence due to blockage of channel AB while blade C receives air at reduced incidence. As a result blade A stalls while blade C may unstall. In this way, stall ‘cell’ shifts along the stage in the direction of lift of the blades. Thus, stall can be stated as a reduction in the lift force at higher angles of incidence.
A rotating stall may rotate in the opposite direction to that of the rotor at about half the rotational speed. It may lead to aerodynamically induced vibrations resulting in fatigue failure of the attached components. Axial flow compressors are more prone to stalling.