Compressor and engine efficiency system and method

A method and system for controlling natural gas compressor and engine units. Particularly, the control method provides for the adjustment of the engine speed and compressor loading, so as to minimize the compressing energy required while maintaining a desired gas flow through the pipeline. The control of the engine speed and compressor loading is based in part on an energy quotient value for the unit. The control method further provides for efficient operation during multiple-stage compression, by controlling the interstage pressure. The control method is also adapted to predict impending compressor and engine unit failures. The control system is based upon a digital-type controller which generates control signals in response to changes in the energy quotient value for the unit.

BACKGROUND AND SUMMARY OF THE INVENTION 
The present invention relates generally to gas engine driven compressor 
stations which are used to transport natural gas. More particularly, the 
present invention concerns a method and apparatus for increasing the 
operating efficiency of natural gas compressor and engine units. 
Until recently, the cost of natural gas did not represent a major portion 
of the daily operating cost for natural gas engine driven compressors such 
as are used in pipe lines to transport natural gas. However, in the past 
few years the cost of fuel gas has increased dramatically. The increased 
cost of natural gas serves to emphasize the necessity to conserve the 
quantity of natural gas consumed by gas driven compressors which transport 
the gas to the market. Improving the efficiency of such gas compressor and 
engine units would not only avoid excessive costs but also a needless 
drain on the limited natural energy resources of the nation. 
Generally speaking, prior art control methods and systems were not 
concerned with maximizing the energy efficiency of a gas compressor and 
engine as a unit. The control methods and systems were directed to such 
concerns as maintaining a desired gas flow, maintaining a certain 
discharge pressure from the compressor, reducing engine vibration, 
preventing an excessive differential pressure across the compressor, and 
maintaining a desired pressure to the consumer. Typically, the speed of 
the engine driving the gas compressor would be set at the rated speed 
supplied by the manufacturer. Then variations in the engine speed would be 
used to effect minor changes in the volume of natural gas being 
transported or to reduce excessive engine vibration. The torque on the 
engine was generally not considered as a control parameter, except for 
preventing an excessive brake mean effective pressure. The compressor load 
setting would be selected on the basis of the characteristic horsepower 
curves for the compressor, so that the available horsepower supplied by 
the engine would be utilized. Examples of prior art methods and systems 
are taught in U.S. Pat. Nos. 4,119,391, A. Rutshtein et al., Oct. 10, 
1978; 3,753,626, R. M. Bacchi, Aug. 21, 1973; 3,716,305, G. Oberlander, 
Feb. 13, 1973; 3,291,378, J. P. Yarnall, Dec. 13, 1966; 3,251,534, H. E. 
Strecker, May 17, 1966. 
In the present invention, the control parameters for the compressor and 
engine are selected so as to minimize the use of natural gas, i.e. the 
compressing energy required, to maintain the desired gas flow through the 
pipeline. Major losses in efficiency arise from unduly light loading of 
the compressor, low torque on the engine, low compression ratio, and/or 
high engine speed. Other factors which affect efficiency are the ambient 
temperature, the suction gas temperature, gas heating value, and the age 
or overhaul date of the engine and compressor. 
The present invention further provides a control method wherein one or more 
gas compressor and engine units can be adapted for multiple stage 
compression operation. These control methods are based upon adjusting 
various control parameters in response to a change in one or more 
operating parameters in a gas compressor and engine unit. The control 
parameters are adjusted so that a maximum energy quotient for the unit is 
achieved. The use of the energy quotient value also allows for precise 
predictions of impending unit failures. 
Additional features and advantages of the present invention will be 
apparent from the following disclosure taken in conjunction with the 
accompanying drawings and appended claims.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENT 
Referring to FIG. 1, a schematic diagram of a typical field or gathering 
gas compressor station 10 is shown. Compressor station 10 and the gas 
compressor and engine units contained therein are described hereinafter to 
illustrate the application of the present invention in the control of like 
compressor stations and units. Compressor station 10 is suitable for use 
in injecting and withdrawing natural gas from two gas fields of a storing 
capacity on the order of 14.5 and 22.0 billion cubic feet of gas. Line 12 
indicates the connecting pipeline for one field, and line 14 indicates the 
connecting pipeline for another field. Before the natural gas from these 
fields reaches the compressor building 16, several events must first 
occur. Scrubbers 18 and 20 remove any liquids or particles that may be 
present in pipelines 12 and 14, by letting them settle to the bottom of 
the scrubbers. Heaters 22 and 24 increase the temperature of the gas to 
prevent freezing the regulators 26. During the winter months when natural 
gas is being withdrawn from the fields, the pressure in the fields will be 
substantially greater than the pressure in the gas transmission lines. For 
example, the maximum pressure limit in these two fields might be 1780 psi; 
whereas the typical gas transmission line pressure is 800 psi. The 
regulators 26 provide the necessary pressure drop before the natural gas 
reaches the transmission line. However, a rapid decrease in pressure also 
acts to cool the gas at a rate of approximately 7.degree. F. per 100 psi 
drop in pressure. Hence, the heaters will be needed to protect the 
regulators from freezing when the pressure in the field is much greater 
than that in the gas transmission line. Interposed between the heaters and 
the regulators are meters 22 and 24. These orifice type meters are used to 
measure the volume of gas flow from the fields. 
The compressor building 16 contains two gas compressor and engine units, as 
indicated by the two sets of pipelines connected to the building. Both the 
suction gas pipeline 28 leading to the compressor building and the 
discharge gas pipeline 30 leaving from the compressor building have a 30 
inch diameter. These two pipelines are connected through a series of 
valves and scrubbers 31 to pipelines 32 and 34, which are used to 
transport natural gas to and from another compressor station. The coolers 
36 in discharge gas line 30 are used to reduce the temperature of the gas 
after compressing when the field is being charged with natural gas. Line 
38 has a 4 inch diameter, and connects the compressor building with the 
blowdown stack 40. The blowdown stack is essentially a vent to atmosphere. 
It is used to purge the compressors of air before starting, and 
de-pressurize the compressors after shutdown. The valves generally 
designated at 31 also provide a direct connection between regulators 26 
and gas transmission pipelines 32 and 34, so that the gas compressor and 
engine units in building 16 may be bypassed. It should be appreciated that 
this bypass connection allows the gas to be injected into the fields, or 
withdrawn from the fields directly from a downstream compressor station. 
In operation, this compressor station charges the fields with natural gas 
during the summer months, and withdraws the gas from the field during the 
winter months. During the charging or injecting cycle, the fields may be 
charged from one or both of the compressor and engine units in building 
16. Typically both compressor and engine units would be connected in 
parallel for single-stage operation until the pressure in the fields would 
reach 1330 psi. Then, under multiple-stage compression, the fields would 
be fully charged to 1780 psi. Typically, the downstream compressor station 
units would provide the first stage, and the units at compressor station 
10 would provide the second stage. It may also be appreciated that the two 
units at compressor station 10 could be adapted to be connected in series 
for multiple-stage compression, rather than utilize the downstream 
compressor station units for the first stage. During the early withdrawal 
cycle, the pressure in the fields will be sufficient to transport the gas 
without resorting to the use of the compressor. Consequently, the valves 
at 31 will be actuated to provide a direct connection between the 
regulators and the gas transmission pipelines. When the pressure in the 
fields is no longer sufficient to transport a desired capacity or volume 
of gas, the units at compressor station 10 would again be utilized. 
Referring to FIG. 2, a plan view of a gas compressor and engine unit 42 is 
shown. The compressor module 44 contains four identical ends 36, which 
each house a double-acting piston. The compressor may be characterized as 
a reciprocating, positive displacement, double-acting mechanism. The 
engine module 48 is a 12 cylinder, single-acting, reciprocating, V-type 
internal combustion engine. This engine operates on natural gas, and is 
rated at 4000 brake horsepower (BHP). FIG. 3 illustrates a side elevation 
view of this gas compressor and engine unit. The overall length of this 
unit is 39 ft., 7 in., the maximum width across the compressor ends is 18 
ft., 8 in., and the maximum height at the engine is 15 ft., 4 in. 
The two gas compressor and engine units at compressor station 10 are 
similar to the unit shown in FIGS. 2 and 3. The primary difference is that 
the engine modules at compressor station 10 contain 8 cylinders, and are 
each rated at 2000 BHP. 
Referring to FIG. 4, a fragmentary cut-a-way prospective view of the 
compressor module 44 is shown. Each compressor cylinder end 46 contains a 
single double-acting piston 50. Piston rod 52 is attached to piston 50 at 
the cylinder head end, and is attached to connecting rod 54 at the crank 
end. Connecting rod 54 is secured to crankshaft 56, which is in turn 
coupled to the engine module 48. A suction gas port 58 is located at the 
top of each compressor cylinder end 46, and an identical port located at 
the bottom of the cylinder ends is used for the gas discharge. Adjacent to 
each suction gas port are four plate and poppet type valves 60 (a, b, c, 
d), which control the flow of suction gas into the compression chamber 61 
by responding to a pressure differential across the valve. The lower set 
of four valves 62 (a, b, c, d) adjacent to the discharge gas port control 
the flow of gas from the compression chamber 61. Cylinder head 64 contains 
a fixed volume pocket 66, which is controlled by handwheel 68. In addition 
to the manual clearance control shown, automatic control may be effected 
through the use of pneumatically actuated valves. 
In operation, the capacity of gas flow (cubic feet/hour) from compressor 
module 44 may be controlled by three mechanisms. First, the speed of the 
engine will of course control the speed at which the pistons 50 
reciprocate in the compression chamber 61. This affects the actual volume 
of gas displaced by the piston as it travels the length of its stroke, 
which is referred to as the piston displacement (PD). Second, the volume 
remaining in the compression chamber at the end of a discharge stroke, 
referred to as the cylinder clearance, may be adjusted by opening and 
closing pockets, such as fixed volume pocket 66. This affects the 
volumetric efficiency (VE), which is the ratio of the compression chamber 
capacity to the actual volume displaced by the piston. Third, the number 
of compressor ends 46 may be varied or an end may be changed to 
single-acting operation, by the use of unloaders (not shown) attached to 
the suction valves 60 (a, b, c, d). The unloaders act to hold the suction 
valves open, and thereby prevent the gas from being discharged from the 
compressor. For example, when piston 50 moves to the left (cylinder head 
end), valves 60c and d would be opened to allow the gas to fill the 
portion of the compression chamber. At the cylinder head end, valves 60a 
and b would be closed to prevent the gas from escaping back into the 
suction gas port, and valves 62a and b would be open to allow the gas to 
be discharged from the cylinder head portion of the compression chamber, 
to the discharge gas port. If an unloader opened either suction valves 60a 
or b, the gas in the compression chamber would preferentially escape back 
into the suction gas port rather than through discharge valves 62a and b, 
due to the greater pressure at the discharge gas port than at the suction 
gas port. In this situation the compressor end would be characterized as 
single acting, as only the portion of the compression chamber nearest the 
crank end would be pumping gas through the compressor. Similarly, if all 
of the suction valves 60 (a, b, c, d) were unloaded, no gas would be 
pumped through the compressor end even though the piston would be 
reciprocating at the same speed. 
As an example of the foregoing, Table 1 illustrates the theoretical volume 
of gas displaced by the pistons as a function of various compressor 
loadings, for one of the gas compressor and engine units at compressor 
station 10. These values were calculated at the rated engine speed (600 
rpm) of the engine module. Although there are four compressor ends 46 in 
compressor module 44, the term "ends out" in Table 1 encompasses single as 
well as double-acting piston operation. In other words, the availability 
of single-acting operation provides in effect eight possible compressor 
ends which may be utilized to pump natural gas through the compressor. As 
indicated in the Table, the gas compressor and engine unit at compressor 
station 10 has been adapted to include one large (L) and two small (S) 
pockets for each compressor end 46. These pockets are attached to the 
compressor cylinder head, and are pneumatically actuated. The large 
pockets have a volume of 1027 cubic inch, and the small pockets each have 
a volume of 300 cubic inch. As stated previously, the number of pockets 
open or closed control the clearance volume left in the compression 
chamber after the piston has completed a compression stroke. As indicated 
in the table, this volume may even exceed the volume of gas displaced by 
the piston stroke. Although the cylinder clearance does not affect the 
volume of gas displaced by the pistons, it does affect the volume of gas 
flowing from the compressor. As will be described later, the piston 
displacement is only one of several terms defining the capacity (Q) of gas 
flow from the compressor. The cylinder clearance also affects the torque 
on the engine. and in combination with the speed of the engine, the 
cylinder clearance provides an effective control of the torque. 
TABLE I 
______________________________________ 
THEORETICAL PISTON DISPLACEMENT AS A 
FUNCTION OF COMPRESSOR LOADING 
COMPRESSOR 
PISTONS AVERAGE 
ENDS POCKETS DISPLACEMENT CYLINDERS 
OUT OPEN CF/HR CLEARANCE % 
______________________________________ 
0 0 134,550 30.67 
0 1-S " 36.32 
0 2-S " 39.98 
0 4-S " 49.29 
0 6-S " 58.60 
0 8-S " 67.91 
0 6S-1L " 74.54 
0 8S-1L " 83.85 
0 6S-2L " 90.48 
0 8S-2L " 99.79 
0 6S-3L " 106.42 
0 8S-3L " 115.73 
0 6S-4L " 122.35 
0 8S-4L " 131.60 
1 0 116,698 32.59 
1 1-S " 37.98 
1 2-S " 43.35 
1 4-S " 54.08 
1 6-S " 64.82 
1 8-S " 75.55 
1 6S-1L " 83.19 
1 8S-1L " 93.93 
1 6S-2L " 101.57 
1 8S-2L " 112.30 
1 6S-3L " 119.95 
1 8S-3L " 130.68 
2 0 98,846 35.25 
2 1-S " 41.58 
2 2-S " 47.92 
2 4-S " 60.59 
2 6-S " 73.27 
2 8-S " 85.95 
2 6S-1L " 94.97 
2 8S-1L " 107.64 
2 6S-2L " 116.66 
2 8S-2L " 129.34 
3 0 80,993 39.05 
3 1-S " 46.78 
3 2-S " 54.52 
3 4-S " 69.98 
3 6-S " 85.45 
3 8-S " 100.92 
3 6S-1L " 111.93 
3 8S-1L " 127.39 
4 0 63,137 45.00 
4 1-S " 54.92 
4 2-S " 64.84 
4 4-S " 84.68 
4 6-S " 104.52 
4 8-S " 124.36 
______________________________________ 
In describing the present invention, a particular nomenclature will be 
utilized. Although this nomenclature is more or less standard to those 
skilled in the art, a glossary providing definitions of the nomenclature 
is set forth in Table 2 for convenience and clarity. 
TABLE 2 
______________________________________ 
GLOSSARY 
______________________________________ 
A.sub.h = area of the piston head (sq. in.) 
A.sub.r = area of the piston rod (sq. in.) 
BHP = brake horsepower 
BTU = British thermal unit 
CF/HR = cubic feet per hour 
C.sub.v = cylinder clearance volume 
E.sub.i = ideal energy required to compress natural 
gas (BTU/SCF) 
E.Q. = energy quotient 
F.sub.h = fuel heating value (BTU/SCF) 
Input = energy needed to operate the engine 
(BTU/HR) 
k = ratio of specific heats 
L = length of piston stroke (in.) 
LHV = lower heating value of fuel gas (BTU/SCF) 
MM-BTU/HR = millions of BTU per hour 
N = engine speed (rpm) 
Output = Q .times. Z.sub.s .times. E.sub.i (BTU/HR) 
overall 
efficiency 
= compressor cylinder efficiency .times. mechanical 
efficiency 
P. D. = piston displacement (CF/HR) 
P.sub.d = discharge gas pressure (PSIG) 
P.sub.s = suction gas pressure (PSIG) 
Q = capacity of gas flow from the compressor 
(SCF/HR) 
R.sub.c = compression ratio = P.sub.d /P.sub.s 
SCF/HR = standard cubic feet per hour 
s.g. = specific gravity 
T = torque 
t.sub.s = suction gas temperature (.degree.F.) 
V.E. = volumetric efficiency 
V.sub.e = volume of fuel gas consumed by engine 
per hour (SCF/HR) 
Z = compressibility of gas 
______________________________________ 
In order to evaluate the performance of a gas compressor and an engine as a 
unit, the concept of an energy quotient (E.Q.) was developed. The energy 
quotient is essentially the thermal efficiency of the gas compressor and 
engine unit, and is generally defined as 
##EQU1## 
The Output is the theoretical energy required to compress a certain volume 
of gas between given pressure limits, and the Input is the energy consumed 
by the engine driving the compressor. 
Before proceeding to set forth the equations defining the Output and Input, 
several terms used in these equations will first be described. The volume 
of gas displaced per hour by a double acting piston is defined by 
EQU P.D. (SCF/HR)=((2A.sub.h -A.sub.r).times.L.times.N)/28.8, (2) 
at standard conditions of 14.7 (PSIG) gas pressure and 60.degree. (F) gas 
temperature. The volumetric efficiency (V.E.) is the ratio of actual 
cylinder or compression chamber volume to piston displacement (P.D.), and 
is defined as 
EQU V.E.=0.98-C.sub.v (R.sub.c.sup.1/k -1), (3) 
where the ratio of specific heats (k) for natural gas is approximately 1.3. 
The compressibility of gas (Z) is a dimensionless factor which varies with 
temperature and pressure. FIG. 5 illustrates a graph of the theoretical 
compressibility of nature gas, based on a specific gravity of 0.6. The 
theoretical volume per hour of gas flow from the compressor, capacity (Q), 
is defined by 
EQU Q(SCF/HR)=P.D..times.V.E..times.(Actual N/Rated N).times.(P.sub.s 
/14.7).times.(1/Z.sub.s). (4) 
The final term necessary to define the Output of the compressor is the 
ideal thermal energy (E.sub.i) required to compress the gas at standard 
temperature and pressure conditions. FIG. 6 illustrates a graph of the 
ideal (frictionless adiabatic) energy required as a function of the 
compression ratio (R.sub.c). This curve was calculated for natural gas 
with a specific gravity of 0.6 and a k of 1.3. The Output may now be 
defined as 
EQU Output (BTU/HR)=Q.times.Z.sub.s .times.E.sub.i, (5) 
where Z.sub.s is the compressibility of the suction gas. The Input is 
defined as 
EQU Input (BTU/HR)=V.sub.e (SCF/HR).times.F.sub.h (BTU/SCF), (6) 
where the fuel heating value (F.sub.h) is assumed to be the lower heating 
value (LHV) of the fuel gas. 
Although the energy quotient provides an excellent criterion by which the 
performance of a gas engine and compressor unit may be evaluated, it does 
not supply all of the information necessary to control the operation of 
the unit in accordance with the present invention. The brake horsepower 
(BHP) required from the engine and the percent torque (T) on the engine 
are also needed; and are calculated as follows. 
##EQU2## 
The overall efficiency is the compressor cylinder efficiency multiplied by 
the mechanical efficiency of the unit. FIG. 7 illustrates a graph of the 
overall efficiency as a function of the compression ratio for a particular 
unit: 
The practical application of the above principals in the control of one or 
more gas compressor and engine units will now be described. One of the 
primary concerns in the operation of a compressor station or unit is the 
amount of gas (volume/hour) being transferred through the pipeline. During 
the time when gas is being stored in a field for future use, the amount of 
gas being injected into the field would normally not be considered 
critical. In fact, this would generally be dependent upon the geological 
formation of the field. However, when gas is being withdrawn from the 
field to meet a required demand by the consumer, the maintenance of a 
constant volume of gas flow from the field is quite important. This is 
especially true in the winter months when natural gas is being used to 
heat many residential homes. Consequently, the situation may arise where 
the station or unit is adjusted to move the gas in the most efficient 
manner, even though the volume of gas transferred is somewhat reduced. 
Further, the situation may arise where the volume of gas being transferred 
is controlling, and the efficiency of the station or unit can only be 
optimized within this constraint. 
In terms of the field compressor station 10 illustrated in FIG. 1, the 
following basic control options exist: one unit may be utilized to 
transfer gas; both units may be combined in parallel; both units may be 
combined in series (multiple-stage compression); the gas pressure in the 
fields may be sufficient to transfer gas from the field without the units; 
or the down stream station may be utilized to transfer gas to or from the 
fields (single or multiple-stage). 
In the situation where one gas compressor and engine unit is utilized and 
the volume of gas flowing from the compressor is not critical, then in 
accordance with the present invention the unit should be operated so that 
the energy quotient (E.Q.) is maximized. This is accomplished through 
adjustments of the speed of the engine (N) and the loading on the 
compressor. As described previously, compressor loading adjustments are 
performed by varying the number of compressor ends 46 being utilized to 
pump the gas, and varying the cylinder clearance volume (C.sub.v). 
As the engine speed and compressor loading are the only two parameters 
which may be directly controlled, they will be referred to as the "control 
parameters". The remaining parameters which may be physically sensed 
during the operation of a unit, will be referred to as the "operating 
parameters". These include the suction gas pressure (P.sub.s), the 
discharge gas pressure (P.sub.d), the suction gas temperature (t.sub.s), 
the capacity of gas flow from the compressor (Q), the volume of fuel gas 
consumed by the engine (V.sub.e), and the lower heating value of the fuel 
gas (LHV). 
The effect of the engine speed (N), the engine torque (T), and the 
compression (R.sub.c) on the energy quotient for a unit is illustrated in 
FIG. 8. This composite graph was taken from experimental data on another 
type of gas compressor and engine unit than that disclosed herein. Each 
curve was based on maintaining the other two parameters (N, T, or R.sub.c) 
constant. It may be observed that the energy quotient decreases when the 
engine speed increases. As the rated speed for this unit is 330 (rpm), it 
is apparent that the engine should be operating at a lower speed to 
maximize the efficiency of the unit. For the purpose of the present 
invention, the effect of the engine speed on the efficiency of the engine 
itself is unimportant, and should be distinguished from the efficiency of 
the compressor and engine as a unit. 
With reference to the torque on the engine, it may be observed that the 
energy quotient increases linearly with an increasing torque. From this 
curve it is apparent that the unit operates most efficiently when the 
torque on the engine is approximately 100% of the rated torque. As stated 
previously, the torque on the engine is controlled by the engine speed and 
the compressor loading. Therefore, a gas compressor and engine unit should 
be controlled so that the compressor loading maximizes the engine torque 
while maintaining the engine speed at the minimum value necessary to pump 
the desired capacity of natural gas through the compressor. 
The third curve in FIG. 8 illustrates the relationship between the 
compression ratio (R.sub.c) and the energy quotient of the unit. As 
indicated, the compression ratio has a substantial effect on the energy 
quotient. However, the compression ratio is one of the last controllable 
parameters in the operation of the unit. For example, when natural gas is 
being withdrawn from the field, the suction gas pressure (P.sub.s) at the 
compressor will be dependent upon the gas pressure in the field as well as 
the capacity of gas being drawn from the field. 
The compression ratio is also important because it is the only sensed 
parameter which changes during the normal operation of a gas compressor 
and engine unit. However, the variation in this one parameter also affects 
the compressibility of the gas (Z), the volumetric efficiency (V.E.), the 
ideal energy required to compress the gas (E.sub.i) and the overall 
efficiency of the unit. Variations in these factors in turn affect the 
capacity of gas flow (Q), and the energy quotient (E.Q.). Consequently, in 
order to operate a unit so that the energy quotient is maximized, the 
compression ratio must be monitored or sampled at determinable intervals. 
Thus, when the compression ratio changes, the control parameters may be 
adjusted in response to this change to either attempt to maintain the 
original energy quotient, or achieve the highest energy quotient under the 
particular circumstances. It should also be noted that when either of the 
control parameters are adjusted, the compression ratio will again be 
changed. Therefore, this adjustment process will typically be an iterative 
one. 
In the situation where a specific capacity of gas flow from the compressor 
must be maintained, the following exemplifies the proper control steps to 
be taken. Assuming that the unit is withdrawing gas from the field and the 
suction gas pressure decreases, then the capacity (Q) will also decrease. 
First, the engine torque (T) must be examined in order to determine if it 
is below the rated torque for the engine. If the torque may be increased, 
then the option exists to increase the engine speed or the compressor 
loading. In accordance with the present invention, it is preferred that 
the loading on the compressor be adjusted before adjusting the speed on 
the engine. After the loading has been increased by a fixed increment, 
such as closing a pocket or adding another compressor end, then the system 
must be allowed time to stabilize. This is because the compression ratio 
will be changed by the increased capacity (Q). After this time period, the 
torque must be determined again. If the torque is still below the rated 
torque, the compressor loading may be increased another step. This process 
is repeated until the desired capacity is obtained, or the compressor is 
at maximum loading. If the desired capacity cannot be achieved at maximum 
loading, then the engine speed may be increased. Again, the increase in 
engine speed should not be such as to increase the engine torque beyond 
the maximum torque for the unit. Where the desired capacity is achieved at 
a compressor loading less than the maximum available, then the compressor 
loading and engine speed should be adjusted so that the torque on the 
engine is maximized and the engine speed is at the minimum value necessary 
to pump the desired capacity. 
In the situation where two gas compressor and engine units are combined in 
parallel, the units may be operated essentially independent of one 
another. However, when the units are combined in series for multiple-stage 
operation, the units are considered together under the present invention. 
Rather than optimize the efficiency of one unit or the other, the units 
are controlled so that the energy efficiency for the sum of both units is 
optimized. Particularly, the inter-stage pressure is controlled, while 
still maintaining an essentially equal capacity of gas flow from each 
unit. By controlling the inter-stage pressure, the compression ratio for 
each unit may be controlled. Thus, where two similar units are utilized, 
the inter-stage pressure would be adjusted so that the compression ratio 
for each unit would be approximately equal. However, where the units are 
not matched, this adjustment would be dependent upon the particular units 
used. For example, one unit may have a relatively low energy quotient at a 
certain compression ratio, whereas another unit would have a higher energy 
quotient at a lower compression ratio. Thus, in this situation the 
inter-stage pressure would be adjusted so that the relatively inefficient 
unit would have a higher compression ratio than the more efficient unit. 
The above control method may also be adapted to predict impending gas 
compressor and engine unit failures. This would be accomplished by 
comparing the current energy quotient value for the unit with a standard 
or base value energy quotient, determined from curves similar to those in 
FIG. 8. When the difference between these energy quotient values exceeds a 
predetermined value, the unit would then be examined for defects. With 
respect to the compressor module, such defects could include worn out 
cylinder rider bands or rings, worn rod packings, or defective suction or 
discharge valves. In the engine module, such defects could be related to 
the engine timing, the spark plugs, exhaust or intake valves, the power 
piston rings, or the turbocharger. One method of determining whether a 
defect exists in the compressor module or in the engine module is to 
calculate the current energy quotient value on the basis of the actual 
capacity (Q) of gas flow from the compressor. One or more elbow meters, 
standard in the art, would be connected to the outlet pipeline from the 
compressor to sense the actual capacity. By comparing the energy quotient 
value based on the calculated capacity with the energy quotient value 
based on the sensed capacity, the general location of the defect may be 
determined. For example, if the difference between these values is small, 
then the problem would be with the engine module, as the compressor would 
be pumping the capacity of gas it should be pumping. 
The above control methods may also be embodied in an automatic controller 
device to achieve and maintain the maximum energy quotient for one or more 
gas compressor and engine units. It may be appreciated by one skilled in 
the art that such a device could be constructed from analog or digital 
circuitry. However, a digital controller based upon a microprocessor unit 
will be described here. FIG. 10 illustrates a block diagram of a 
microprocessor based controller 70 according to the present invention. The 
central processing unit 72 may be of a type standard in the industry, such 
as the Zilog Z80 microprocessor chip. The programmable read-only-memory 74 
will contain the program for directing the operation of the controller 
device 70. The random access memory 76 would be used to store the various 
parameter values being sensed by input transducers 78, and store the 
results of the calculations incident to the control of the gas compressor 
and engine unit. Input interface 80 is used to receive the signals from 
input transducers 78 and system control center 82 for signal processing 
before being sent to parallel input-output port 84. Such signal processing 
would generally analog to digital conversion and digital signal 
multiplexing. Examples of typical interfacing schemes standard in the art 
may be found in Automated Process Control Systems: Concepts and Hardware, 
Prentice-Hall, Inc., R. P. Hunter, 1978. Input-output port 84 is also used 
to transmit signals to output interface 86, which essentially provides the 
reverse function of input interface 80. The signals from output interface 
86 may then be sent to system control center 82, or to control devices 88. 
System control center 82 is used to provide operator access to controller 
device 70, and would include a keyboard, a printer, and a cathode ray tube 
display. Control devices 88 would be used to control the engine speed, 
compressor pocket clearance, and compressor pneumatic-type unloader 
valves. Oscillator/clock 90 is used to provide the timing signals 
necessary to operate central processing unit 72. Power supply 92 is used 
to provide the electrical power needed to operate control processing unit 
72, memories 74 and 76, clock 90, input-output port 84, and interfaces 80 
and 86. 
In operation, control commands such as to initiate the operation of the gas 
compressor and engine unit may be entered into controller device 70 via 
the keyboard in system control center 82. Such commands may include the 
specification of a desired capacity (Q), percent rated torque, or energy 
quotient; or a range thereof within which the controller may operate. The 
central processing unit would then compare the capacity, torque, and 
energy quotient values with the desired values stored in memory 76. 
Control signals would then be generated and sent to interface 86, where 
the control parameters would be adjusted in response to the above 
comparison. 
While it will be apparent that the preferred embodiments of the invention 
disclosed are well calculated to fulfill the objects above stated, it will 
be appreciated that the invention is susceptible to modification, 
variation and change without departing from the proper scope or fair 
meaning of the subjoined claims.