Rotary kinetic fluid motors or pumps – Method of operation
Reexamination Certificate
2000-01-06
2002-04-02
Verdier, Christopher (Department: 3745)
Rotary kinetic fluid motors or pumps
Method of operation
C415S017000, C415S029000, C415S049000, C060S039240
Reexamination Certificate
active
06364602
ABSTRACT:
BACKGROUND OF THE INVENTION
The present invention relates to gas turbine control and, in particular to the management of compressor Operating Limit Line using an air-flow measurement technique.
The global market for efficient power generation equipment has been expanding in recent years and is anticipated to continue to expand in the future. The gas turbine combined cycle power plant, consisting of a gas turbine based topping cycle and a Rankine based bottoming cycle, continues to be a preferred choice for power generation due to relatively low plant investment costs and continuously improving operating efficiency of the gas turbine-based combined cycle, which minimizes electricity production costs.
By way of background and with reference to the schematic illustration of
FIG. 1
, a typical combined cycle gas turbine includes, in serial-flow relationship, an air intake or inlet, a compressor, a combustor, a turbine, a heat recovery steam generator (HRSG) and its associated steam turbine. Thus, air enters the axial flow compressor at 10 at ambient conditions. Ambient conditions vary from one location to another and day to day. Therefore, for comparative purposes standard conditions are used by the gas turbine industry. Those standard conditions are 59° F. (15° C.), 14.696 psia (1.013 bar), and 60% relative humidity. The standard conditions were established by the International Standards Organization (“ISO”) and are generally referred to as ISO conditions.
The compressed air enters the combustion system at
12
where fuel is injected and combustion occurs. The combustion mixture leaves the combustion system and enters the turbine at
14
. In the turbine section, energy of the hot gases is converted into work. This conversion takes place in two steps. The hot gases are expanded and the portion of the thermo-energy is converted into kinetic energy in the nozzle section of the turbine. Then, in the bucket section of the turbine a portion of the kinetic energy is transferred to the rotating buckets and converted to work. A portion of the work developed by the turbine is used to drive the compressor whereas the remainder is available for generating power. The exhaust gas leaves the turbine at
16
and flows to the HRSG.
The Brayton cycle is the thermodynamic cycle upon which all gas turbines operate. Every Brayton cycle can be characterized by pressure ratio and firing temperature. The pressure ratio of the cycle is the compressor discharge pressure at
12
divided by the compressor inlet pressure at
10
. The General Electric Co. (GE), and we, define the firing temperature as the mass-flow mean total temperature at the stage
1
nozzle trailing edge plane. Another method of determining firing temperature is defined in ISO document 2314 “Gas Turbine-Acceptance Test”. The firing temperature in that case is a reference turbine inlet temperature and not generally a temperature that exists in a gas turbine cycle; it is calculated using parameters obtained in a field test. Thus, this ISO reference temperature is always less than the true firing temperature as defined by GE, above.
A Brayton cycle may be evaluated using such parameters as pressure, temperature, specific heat, efficiency factors, and the adiabatic compression exponent. If such an analysis is applied to a Brayton cycle, the results can be displayed as a plot of cycle efficiency versus specific output of the cycle. Output per pound of air-flow is an important determination since the higher this value, the smaller the gas turbine required for the same output power. Thermal efficiency is important because it directly affects the operating fuel costs.
Many factors affect gas turbine performance. Air temperature, for example, is an important factor in gas turbine performance. Since the gas turbine receives ambient air as inlet air, its performance will be changed by anything that affects the mass flow of the air intake to the compressor; that is changes from the reference conditions of 59° F. and 14.696 psia. Each turbine model has its own temperature-effect curve as it depends on the cycle parameters and component efficiencies as well as air mass flow.
It is also well known that elevated firing temperature in the gas turbine is a key element in providing higher output per unit mass flow, enabling increased combined cycle efficiency, and that for a given firing temperature, there is an optimal cycle pressure ratio which maximizes combined cycle efficiency. The optimal cycle pressure ratio can be theoretically shown to trend ever-higher with increasing firing temperature. Compressors for these turbines are thus subjected to demands for higher levels of pressure ratio, with the simultaneous goals of minimal parts count, operational simplicity, and low overall cost. Moreover, the compressor must enable this heightened level of cycle pressure ratio at a compression efficiency that augments the overall cycle efficiency. Finally, the compressor must perform In an aerodynamically and aeromechanically stable manner under a wide range of mass flow rates associated with varying power output characteristics of combined cycle operation.
Air consumed by industrial gas turbine engines always contains an unknown amount of airborne, solid and liquid particulate. These include dirt, dust, pollen, insects, oil, sea-water salt, soot, unburned hydrocarbons, etc. Deposits form on the compressor turbo machinery blading when these airborne materials adhere to the blading and to each other, leading to changes in the blade aerodynamic profile, blade surface conditions, and flow incidence angle. This fouling causes a concomitant deterioration in the performance parameters of mass flow, thermodynamic efficiency, pressure ratio and surge pressure ratio. This later influence can cause a degradation in the margin between the operating pressure ratio and the surge line, commonly referred to as the surge margin. Tarabrin et al. advise that the magnitude of performance degradation due to fouling has been noted to be about 5% in mass flow, 2.5% in efficiency, and 10% in output. Moreover, a 5% decrease in mass flow has been associated with a reduction in output by 13% and an increase in heat rate by 5.5%. See Tarabrin et al., “An Analysis of Axial Compressor Fouling and a Blade Cleaning Method,”
Journal of Turbomachinery
, Volume 120, April 1998, Pages 256-261.
The maximum pressure ratio that the compressor can deliver in continuous duty is commonly defined in terms of a margin from the surge pressure ratio line. Compressor surge is the low frequency oscillation of flow where the flow separates from the blades and reverses flow direction through the machine, i.e., it serves as a physical limit to compressor operation at a given speed.
The conventional approach to compressor protection is to program into the gas turbine control a so-called Operating Limit Line that affords a margin, typically between 5 and 25%, from a new and clean compressor surge boundary. One of the considerations in establishing this margin is a fixed allowance for the anticipated level of compressor fouling and the corresponding effect on surge margin. Once set, this allowance is not modified over time and/or operating conditions.
BRIEF SUMMARY OF THE INVENTION
The present invention was derived from the simultaneous need for high cycle pressure ratio commensurate with high efficiency and ample surge margin through-out the operating range of the compressor. The invention provides a design and operational strategy that provides optimal pressure ratio and surge margin for both the case where the inlet guide vanes are tracking along the nominal, full-flow schedule and where the inlet guide vanes (IGVs) are closed-down for reduced flow under Power-Turn-Down conditions.
More specifically, the invention provides for active management of the compressor Operating Limit Line using a flow sensing system to determine the amount of air-flow going through the system. By determining air-flow, and comparing it to an air-flow value stored in the gas turbine control, the degradation of flow with compressor fo
Andrew Philip L.
Cotroneo Joseph A.
Stampfli John David
Yeung Chung-hei
General Electric Company
Nixon & Vanderhye PC
Verdier Christopher
Woo Richard
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